Free turbine type gas turbine engine with variable free turbine guide vane control system

ABSTRACT

A gas turbine engine and method and control therefor particularly useful as the power plant for a ground vehicle. The gas turbine engine is of the free turbine type, and employs variable guide vanes to control the free turbine output. The guide vane control system is a combined hydromechanical and electronic control system to automatically regulate the guide vanes for both power regulation and braking modes.

CROSS-REFERENCE TO RELATED APPLICATIONS

Cross reference is made to the following co-pending U.S. Patentapplications, having common assignee with the present application, allfiled concurrently herewith and entitled "Gas Turbine Engine," and whichrelate to subject matter common with that disclosed in the presentapplication: Ser. No. 863,361 of Woodhouse, et al; Ser. No. 863,570 ofJansen, et al; Ser. No. 863,495 of Hatch, et al; Ser. No. 863,375 ofMattson, et al; Ser. No. 863,365 of Bolliger, et al; Ser. No. 863,198 ofWiher, et al; Ser. No. 863,205 of Sumegi, et al.

BACKGROUND OF THE INVENTION

This invention relates to gas turbine engines, and relates moreparticularly to an improved gas turbine engine and method and controltherefor particularly useful as the power plant for a ground vehicle.

Recent advances in gas turbine engine technology have improved theiroverall efficiency and economy to such an extent that this type of powerplant has become competitive in many instances with more conventionalinternal combustion type power plants such as Otto or Diesel cycleengines. For instance, gas turbine technology has made significantinroads as the power plant for aircraft engines. Similarly, attemptshave been made to develop a gas turbine engine which would becompetitive with the more conventional internal combustion engines inhigh-production ground vehicles such as on-the-road automobiles andheavy trucks. The gas turbine offers significant advantages ofequivalent or better operational efficiency, fuel savings, and lessemissions as well as being able to utilize a variety of different fuelson an economic basis. Further, the gas turbine engine in many instancesoffers greater overall economy over the entire operational life of avehicle.

The inherent operational characteristics of a gas turbine enginepresent, however, certain problems when utilized in a ground vehicle.More specifically, a gas turbine engine generally includes a gasgenerator section which provides a large pressurized air flow to acombustor wherein the air flow is mixed and ignited with fuel to greatlyincrease the temperature of the resulting gas flow. Hot pressurized gasflow then drives one or more turbines to produce useful rotarymechanical output power. Normally one of these turbines is a portion ofthe gas generator section for driving the fan which provides the highvolume pressurized air inlet flow. Downstream power output turbines thengenerate the useful mechanical power output. Conventionally, the highspeed, high volume gas flow from the gas generator drives the turbinesat relatively high speeds. Other inherent characteristics of such gasturbine engines relates to the thermodynamic and aerodynamic processescarried out therewithin which dictate that operational efficiency of theengine increases substantially with increasing maximum temperature ofthe gas flow.

These operating characteristics of a gas turbine engine present certaindisadvantages in comparison to the normal operation of reciprocating orrotary piston type internal combustion engines for ground vehicles. Moreparticularly, the internal combustion engine inherently provides asubstantial amount of deceleration horsepower for the vehicle uponreducing fuel flow thereto through the drag imposed by the reciprocatingportion of the engine. In contrast, the high rotational inertia of theturbines of the gas turbine engine normally do not permit suchimmediate, relatively high horsepower braking for a ground vehiclesimply upon reducing fuel flow to the combustor of the gas turbineengine. To overcome this disadvantage, a variety of proposals have beenoffered in the past to increase the braking characteristics of a gasturbine engine when utilized for driving a ground vehicle. Primarily,these concepts relate to completely extinguishing the combustion processwithin the combustor to produce maximum dynamic braking. However,operational life of a gas turbine engine is substantially reduced bycontinual thermal cycling of the entire engine as created uponextinguishing the combustion process. Further, such approaches adverselyaffect emissions. Other concepts relating to improving the dynamicbraking characteristics of a gas turbine engine revolve around theutilization of a "fixed shaft" type of gas turbine engine wherein thegas generator section and the power drive section are mechanicallyinterconnected to drive the vehicle. While such an arrangement improvesthe dynamic braking, it greatly reduces the adaptability of the engineto perform various other processes for driving a ground vehicle, and dueto this limited adaptability has met with limited success in use as thepower source for a high-production type of ground vehicle. An example ofsuch prior art structure is found in U.S. Pat. No. 3,237,404. The normalmethod for dynamic braking in gas turbine powered aircraft, thrustreversal, is of course not readily applicable to ground vehicles.

Prior arrangements for gas turbine engines for ground vehicles also havesuffered from the disadvantage of not providing efficient, yet highlyresponsive acceleration in comparison to internal combustion engines.Inherently, a free turbine engine normally requires a substantiallylonger time in developing the maximum torque required duringacceleration of the ground vehicle. Prior attempts to solve this problemhave centered about methods such as operating the gas generator at aconstant, maximum speed, or other techniques which are equallyinefficient in utilization of fuel. Overall, prior gas turbine enginesfor ground vehicles normally have suffered from a reduced operationalefficiency in attempting to improve the acceleration or decelerationcharacteristics of the engine, and or resulted in reduced efficiency bysubstantially varying the turbine inlet temperature of the gas turbineengine which is a primary factor in the fuel consumption of the engine.Further, prior art attempts have generally been deficient in providing areliable type of control system which is effective throughout alloperational modes of a gas turbine engine when operating a groundvehicle to produce safe, reliable, operating characteristics. Further,such prior art gas turbine engines have resulted in control arrangementswhich present a substantial change in required operator actions incomparison to driving an internal combustion powered vehicle.

Other problems related to prior art attempts to produce a gas turbineengine for ground vehicle relate to the safety and reliability of thecontrol system in various failure modes, safe and reliable types ofcontrols, and in the overall operational efficiency of the engine. Amajority of these problems may be considered as an outgrowth of attemptsto provide a gas turbine engine presenting operational characteristicsduplicative of the desirable, inherent actions of an internal combustionengine.

Accordingly, it will be seen that it would be highly desirable toprovide a gas turbine engine and associated controls which incorporatethe desirable operational features of both a gas turbine and internalcombustion engine, but while providing an economical end product ofsufficiently reliable and safe design for high volume production basisfor ground vehicles.

Discussions of exemplary prior art structure relating to the engine ofthe present invention may be found in U.S. Pat. Nos. 3,237,404 discussedabove; 3,660,976; 3,899,877; 3,941,015 all of which appear to relate toschemes for transmitting motive power from the gas generator to theengine output shaft, and 3,688,605; 3,771,916 and 3,938,321 that relateto other concepts for vehicular gas turbine engines. Examples ofconcepts for variable nozzle engines may also be found in U.S. Pat. Nos.3,686,860; 3,780,527 and 3,777,479. Prior art fuel governor controls inthe general class of that contemplated by the present invention may befound in U.S. Pat. Nos. 3,400,535; 3,508,395; 3,568,439; 3,712,055;3,777,480 and 3,913,316, none of which incorporate reset and overridefeatures as contemplated by the present invention; and 3,521,446 whichdiscloses a substantially more complex fuel reset feature than that ofthe present invention. Examples of other fuel controls less pertinent tothe present invention may be found in U.S. Pat. Nos. 3,851,464 and3,888,078. U.S. Pat. No. 3,733,815 relates to the automatic idle resetfeature of the present invention while U.S. Pat. Nos. 2,976,683;3,183,667 and 3,820,323 relate to the scheduling valve controls.

SUMMARY OF THE INVENTION

An important object of the present invention is to provide an improvedgas turbine engine and method and more particularly arrangementsexhibiting desirable operational features normally inherent to pistonengines.

Another important object is to provide provisions producing improvedfuel performance in a variety of operations of a ground vehicle drivenby a gas turbine engine.

Another important object of the present invention is to provide improvedacceleration, deceleration characteristics for a gas turbine drivenground vehicle, and to provide a more reliable, longer life gas turbineengine for propulsion or power generating purposes.

In summary, the invention contemplates a recuperated, free turbine typeengine with separate gas generator and power turbine sections. A fuelgovernor controls fuel flow to the combustor to set gas generator speedin relation to the throttle lever. Reset solenoids can override andadjust fuel flow in response to certain operating parameters orconditions of engine operation. For instance, in response to low speedon the output shaft of the drive train clutch which is indicative of animpending desired engine acceleration for increased torque output, areset solenoid increases fuel flow and the gas generator idle speed tosubstantially reduce time required in increasing engine torque output. Ascheduling valve is effective to control fuel flow during engineacceleration to prevent excessive recuperator inlet temperature andmaintain turbine inlet temperature at a substantially constant, highlevel for maximum engine performance. The scheduling valve is responsiveto combustor inlet gauge pressure and temperature, and also controlsfuel flow during deceleration in a manner maintaining combustion.Variable turbine guide vanes are shifted first to maximize powerdelivered to the gas generator during its acceleration, and subsequentlyare shifted toward a position delivering maximum power to the powerturbine section. The variable guide vane control includes ahydromechanical portion capable of controlling power turbine sectionspeed in relation to throttle position, and has an electromechanicalportion cooperable therewith to place the guide vanes in a braking modefor deceleration. Power feedback is incorporated to provide yet greaterbraking characteristics. When such is selected, the gas generator speedis automatically adjusted to approach power turbine speed, then througha relatively low power rated clutch the gas generator and power turbinesections are mechanically interconnected such that the rotationalinertia of the gas generator section assists in retarding the engineoutput shaft.

These and other objects and advantages of the present invention are setforth in or will become apparent from the following detailed descriptionof a preferred embodiment when read in conjunction with the accompanyingdrawings.

BRIEF DESCRIPTION OF THE DRAWINGS

In the drawings:

FIG. 1 is a left front perspective illustration of a gas turbine engineand associated drive train embodying the principles of the presentinvention;

FIG. 2 is a perspective illustration of the power feedback drive trainas incorporated in the engine with portions of the engine shown inoutline form;

FIG. 3 is a fragmentary, partially schematic, elevational cross-sectionof the power feedback clutch and associated hydraulic system, takengenerally along lines 3--3 of FIG. 2;

FIG. 4 is a partially schematic cross-sectional representation of therotating group of the engine with controls associated therewith shown inschematic, block diagram form;

FIG. 5 is a right front perspective view of a portion of the housing,ducting passages and combustor of the engine with portions broken awayto reveal internal details of construction;

FIG. 6 is a partially schematic, plan cross-sectional view of the fuelgovernor 60 with portions shown perspectively for better clarity ofoperational interrelationships;

FIG. 6a is an enlarged partial elevational cross-sectional view of thefuel pump taken generally along lines 6a--6a of FIG. 6;

FIGS. 6b, 6c, 6d are enlarged cross-sectional views of a portion of thefuel governor control showing different operational positions ofsolenoid 257;

FIG. 7 is a schematic, cross-sectional and perspective functionalrepresentation of scheduling valve 62;

FIG. 8 is a plan cross-sectional view through one portion of thescheduling valve;

FIG. 9 is a plan cross-sectional view of the scheduling valve takengenerally along lines 9--9 of FIG. 8;

FIGS. 10 and 11 are enlarged views of portions of valve 282 showing theinterrelationship of fuel metering passages as would be viewedrespectively along lines 10--10 and 11--11 of FIG. 7;

FIG. 12 is a schematic cross-sectional representation of guide vanecontrol 66;

FIG. 13 is an exploded perspective illustration of the guide vanes andactuator linkage;

FIGS. 14, 15 and 16 are circumferential views showing variousoperational relationships between the variable guide vanes and the powerturbine blades;

FIG. 17 is a schematic logic representation of a portion of theelectronic control module 68;

FIG. 18 is a graphical representation of the area ratio across the powerturbines as a function of guide vane angle;

FIG. 19 is a graphical representation of the desired gas generatorsection and power turbine section speeds selected in relation tothrottle position; and

FIG. 20 is a graphical representation of the relationship of fuel flowpermitted by the scheduling valve as a function of combustor pressurealong lines of constant combustor inlet temperature.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

With reference to the figures, listed below are the abbreviationsutilized in the following detailed description to denote variousparameters:

N_(pt) =Power Turbine 54 Speed

N_(gg) =Gas Generator 52 Speed

N_(gg*) =Preselected Gas Generator 52 Speed

N_(ti) =Transmission Input Shaft 36 Speed

e=Predetermined Minimum Speed of Transmission Input Shaft 36

W_(f) =Fuel flow

B=Stator Vane 120, 122 Angle

B*=Predetermined Stator Vane Angle

a=Throttle 184 Position

a*=Predetermined Throttle Position

T₂ =Compressor Inlet Temperature

P₂ =Ambient Pressure

T₃.5 =Combustor Inlet Temperature

P₃.5 =Combustor Pressure

P₃.5* =Preselected Intermediate Value of Combustor Pressure

T₄ =Turbine Inlet Temperature

T₆ =Turbine Exhaust Temperature

Engine 30

Referring now more particularly to the drawings, an improved gas turbineengine as contemplated by the present invention is generally denoted bythe numeral 30. As depicted in FIG. 1 the engine is coupled to asubstantially standard drive train for a vehicle, particularly a truckin the 450 to 600 horsepower class, with a power output shaft 32 as theinput to a drive train clutch 34. A transmission input shaft 36 extendsbetween the clutch 34 and a "change speed" type of transmission 38.Transmission 38 is of the manually shiftable gear type; however, it isto be understood that various improvements of the present invention areequally usable with other types of speed varying transmissions. As isconventional the transmission 38 has a variety of different positionsincluding several forward gears, reverse gearing, and a neutralposition. In the neutral position no power is transmitted between thetransmission input shaft 36 and the transmission output shaft 40 whichconventionally extends to the final drive 42 and drive wheels 44 of thevehicle. A manual shifting lever 46 provides selection of the desiredgear ratio, and a speed sensor 48 generates a signal indicative of thespeed of transmission input shaft 36. As schematically depicted in FIG.1 and described in greater detail hereinbelow, the speed sensor 48 maybe of any type compatible with the control medium of the engine 30.Preferably, speed sensor 48 generates an electrical signal transmittedby conductor 50 to the electronic control module of the engine.

Referring to FIGS. 1-4, engine 30 is of the free turbine, recuperatedtype incorporating a gas generator section 52, a power turbine 54mounted on a shaft separate from that of the gas generator 52, and arecuperator 56 that scavanges waste heat from the exhaust flow from theengine for preheating the compressed fluid prior to the combustionprocess. The engine further generally includes a source 58 ofcombustible fuel, a fuel governor generally denoted by numeral 60 whichalso includes the fuel pump therein, a scheduling valve 62 forcontrolling fuel flow normally during acceleration or deceleration ofthe engine through a fuel line 64 extending to the gas generator section52, and a control 66 for variably positioning variable stator vanesincluded in the power turbine section 54. An electronic control module68 receives and processes various input parameter signals and producesoutput control signals to the governor 60 and vane actuator control 66.

Conventionally, there is included an electrical storage battery 70 andassociated starter motor 72 which is preferably selectively coupled toboth the gas generator 52 and a starter air pump 74. During startingoperation, the motor 72 is energized to drive both an air starter pump74 as well as the main gas generator shaft 76. As clearly illustrated inFIG. 2, the preferred form of the invention also includes a drive train78 associated with gas generator shaft 76, and another drive train 80associated with and driven by a main shaft 82 of the power turbine 54.The two drive trains 78 and 80 are selectively interengageable through arelatively low power, wet clutch generally denoted by the numeral 84.This clutch is generally referred to as the power feedback clutch andthe structure thereof is described in detail below with respect to FIG.3, while its functional operation is described further below with regardto the power feedback operation of the present invention.

Gas generator 52 generally includes an appropriately filtered air inlet86 through which ambient air is supplied to a pair of series arrangedcentrifugal compressors 88 and 90. Cross-over ducting 92 carries thecompressed air flow from the first compressor 88 to the secondcompressor 90. The gas generator 52 further includes ducting 94 asdepicted in FIG. 5 which surrounds and collects the compressed air flowexhaust from the circular periphery of the second stage compressor 90,and carries this compressed air flow in a pair of feeder ducts 95 torecuperator 56 in non-mixing, heat exchange relationship with therecuperator. While various forms of recuperator structure may beutilized in conjunction with the present invention, an exemplary form isas described in U.S. Pat. No. 3,894,581 entitled "Method of ManifoldConstruction for Formed Tube-Sheet Heat Exchanger", dated July 15, 1975,issued to Fred W. Jacobsen et al. Though not necessary to theunderstanding of the present invention, reference may be made to theabove referenced patent for a detailed description of a recuperator andits operation. For purposes of the present invention, it is sufficientto state that the compressed air flow from ducts 95 is preheated in therecuperator by the waste heat from the exhaust flow from the engine. Thepreheated, compressed air flow is then ducted through duct 96 to acan-type combustor 98. As best seen in FIG. 5, heated flow from therecuperator passes through a plurality of openings 97 into a plenumportion of duct 96, then through openings 97-a in a portion of thehousing structure supporting combustor 98. Combustor 98 has a perforatedinner liner 99, and airflow from openings 97-a passes into the zonebetween the inner and outer liner to then pass through the perforatedinner liner 99 into the combustor zone. One or more electrical ignitionplugs 100 are suitably connected to a source of high voltage electricalenergy in a conventional manner. The igniter plug is operable tomaintain a continuous combustion process within the interior of thecombustor wherein the fuel delivered from line 64 is mixed and burnedwith the compressed air flow from duct 96.

The gas generator 52 further includes a gas generator turbine 102 of theradial inflow type. The compressed, heated gas flow from combustor 98 isdelivered across turbine inlet choke nozzles 104 disposed in a circulararray about the annularly shaped inlet 106 to the gas generator turbinesection. During engine operation, nozzles 104 maintain pressure incombustor 98 at a level higher than ambient. Flow of this heated,compressed gas across turbine 102 causes high speed rotation of theturbine and the gas generator main shaft 76. This rotation of coursedrives the two centrifugal compressors 88 and 90. Shaft 76 isappropriately mounted by bearings 108 to the stationary housing 110 ofthe engine.

Power turbine section 54 generally includes a duct section 112 andappropriate vanes 114 therein for directing the flow of gases from thegas generator power turbine 102 toward a pair of axial power turbines116 and 118 mounted to the power turbine main shaft 82. The powerturbine section further includes sets 120 and 122 of variablypositionable guide vanes respectively disposed upstream of associatedaxial turbines 116, 118 and their associated blades 117, 119. Asdepicted in FIG. 13, each of the sets of variable guide vanes 120 and122 are disposed in an annular array within the gas flow path and areboth mounted to a common actuating mechanism generally referred to bythe numeral 124. The actuating mechanism 124 comprises a pair of ringgears 126 and 128, one of each set of variable vanes, a link 129 affixedto ring gear 126 and secured to ring gear 128 via plate 129-a. Pivotallymounted to the housing is a bell crank 130, and a twisted link 131 hasopposite ends pivotally attached to link 129 and one arm of bell crank130. A linearly shiftable input shaft 368 acts through a pivot link 132and another arm of the bell crank to cause rotation of crank 130 aboutits axis 133 and consequent simultaneous rotation of both ring gears126, 127. Rotation of input shaft 368 rotates each of the ring gears126, 128 about an axis coincident with the rotational axis of powerdriven shaft 82 to cause rotation of the two sets of guide vanes inunison to various positions relative to the direction of gas flowpassing thereby. As shown in FIGS. 14-16, guide vanes 120 are positionedin a central or "neutral" position of FIG. 14 causing substantiallymaximum area ratio and minimum pressure ratio across the downstreampower turbine wheel blades 117 of wheel 116 in order to minimize theamount of power transferred by the gas flow into rotation of the turbine116. The FIG. 14 position is graphically illustrated by the positionarbitrarily denoted 0° in FIG. 18. The guide vanes 120 are variablypositioned toward the FIG. 15 position, noted as the +20° position inFIG. 18, wherein high pressure ratio exists across blades 117 andmaximum power is transmitted from the gas flow to turbine 116 to rotatethe latter and transmit maximum power to shaft 82. Also, the vanes areoppositely rotatable to the FIG. 16 position, noted as the -95° positionof FIG. 18, wherein the gas flow is directed by the variable vanes 120to oppose and tend to retard the rotation of wheel 116. While only vanes120 and blades 117 are illustrated in FIGS. 14-16, it will be understoodby those skilled in the art that substantially identical operationalrelationships exist between vanes 122 and turbine blades 119 of turbine118.

The gas flow upon exiting the last axial turbine 118 is collected in anexhaust duct 134 which leads to the recuperator 56. The power turbineoutput shaft 82 is a part of or operably connected with the power outputshaft 32 of the engine through appropriate speed reduction gearing. Anair or water cooler 87 is also included to cool the lubricating fluid inengine 30 and communicates with fluid reservoir 89 through hose 91.

Fuel Governor 60

Referring now more particularly to FIGS. 4, 6, 6A-6D, the fuel governor60 receives fuel from source 58 through an appropriate filter 136 intoan inlet port 138 of a fuel pump housing 140. It will be apparent tothose skilled in the art that the housing 140 is attached to and may beintegrally formed with another portion of the main engine housing 110.The governor is operable to schedule fuel flow output through either orboth of the output ducts 142, 144 for delivery to the scheduling valve62. The governor 60 is hydromechanical in nature but capable of beingresponsive to externally applied mechanical and electrical signals, andincludes an appropriate drive connection schematically illustrated byline 146, and associated speed reducing gearing 148 as necessary todrive a gear 150 and drive shaft 152. Shaft 152 drives a fuel pump inthe form of a positive displacement rotary gear pump 154 which receivesfuel from inlet port 138 and displaces it at a substantially higherpressure through an output conduit 156. As clearly illustrated in FIG.6A, the gear pump comprises a pair of intermeshing geats 158 and 160,one of which is driven by drive shaft 152 and the other of which ismounted to an idler shaft 162 journaled within housing 140. Supplied inparallel flow arrangement from output conduit 156 are three passages,i.e. output duct 142, bypass bore 164, and main flow metering passage166. Contained in bypass bore 164 is a bypass regulating valve poppet168 slidable within bore 164 to variably meter excess flow from outputconduit 156 to a return passage 170 connected back to the fuel inletport 138. Pressure of fuel in bore 164 urges poppet 168 downwardly toincrease bypass flow through passage 170, which a helical coilcompression spring means 172 acts against the pressure of fuel to urgepoppet 168 upwardly to reduce volume of flow from bore 164 to passage170. Through a pressure passage 182 the lower end of bypass bore 164communicates with fuel supply conduit 64. Thus, pressure of fluid inconduit 64 is exerted upon the lower side of bypass valve poppet 168 toassist spring 172 in opposing the force created by the high pressurefluid in poppet conduit 156. Passage 166 terminates in a metering nozzle174 secured by plate 176 to the housing, and having a reduced diameteropening 178 communicating with a central cavity 180.

The fuel governor 60 further includes a manual throttle input in theform of a throttle lever 184 shiftable between opposed adjustable stops186, 188 adjustably secured to housing 140. Through an appropriatebearing 190 a shaft 192 extending within internal cavity 180 isrotatable relative to housing 140. Integrally carried by shaft 192 in anopen-sided camming section 194 into which are pressed fit a pair of stubshafts 196 that respectively carry rollers 198. Rollers 198 areengageable with the lower shoulder of a spring stop 200 such thatrotation of the throttle lever 184 and shaft 192 causes consequentrotation of stub shafts 196 which are non-aligned with the mainrotational axis of shaft 192, and thus vertical shifting of spring stop200 through rollers 198. During its vertical or longitudinal shifting,spring stop 200 is guided by a guide shaft 202 which has an upper guideroll pin 204 slidably extending through a central bore of spring stop200. Guide rod 202 is threadably received and secured such as by locknut 206 to housing 140.

The governor 60 further includes a mechanical speed sensor whichincludes a flyweight carrier 208 rigidly secured to rotate with shaft152. Rotating with carrier 208 are a plurality of regularly spacedflyweights 210 mounted for pivotal movement upon pins 212 securing theweights 210 to carrier 208. Dependent upon the speed of shaft 152, thecentrifugal force causes rotation of weights 210 about pins 212 to causethe inner ends thereof to shift downwardly as viewed in FIG. 6 and drivethe inner rotating race 214 of a roller bearing assembly alsodownwardly. Through ball bearings 216 this downward force is transmittedto the non-rotating outer race 218 of the bearing assembly to causedownward shifting of non-rotating segment 220. At its lower end segment220 carries a spring stop shoulder 222, and a speeder spring 224operably extends between the stop 222 of segment 220 and the spring stop200 associated with the throttle input mechanism. Through a preload ofspring 224 acting on segment 220 the flyweights are normally urgedupward to the zero or low speed position illustrated in FIG. 6.Increasing speed of shaft 152 causes downward shifting of segment 220.Thus it will be apparent that throttle lever 184 acts essentially toselect gas generator speed as reflected by the speed of shaft 152, sincethe compression of spring 224 is set by rotation of throttle lever 184and then opposed by the centrifugal force created by the rotation ofshaft 152. The vertical position of segment 220 therefore becomesindicative of the difference between selected speed (position of inputthrottle 184) and actual gas generator speed as sensed throughflyweights 210. FIG. 19 illustrates the action of spring 224 inrequesting different levels of gas generator speed N_(gg), as thethrottle is moved through different positions, a.

Governor 60 further includes a main fuel throttle lever 226 pivotallymounted by pin 228 to housing 140. One arm 230 of lever 226 terminatesin a spherically shaped end 230 within a receiving groove 232 on segment220 of the speed error signal mechanism. An opposite arm 234 of lever226 is movable toward and away from metering orifice 178 in response toshifting of segment 220 to thereby variably meter fuel flow from passage166 into internal cavity 180. It will be apparent that the regulatingvalve poppet 168 is variably positioned in response to the pressuredifferential between passage 168 and conduit 64 downstream of themetering orifice 178 to variably meter bypass fluid flow through passage170 in order to maintain a substantially constant pressure differentialacross the fluid metering orifice created between metering opening 178and the arm 234 of fuel lever 226. Thus the rate of fuel flow deliveredfrom passage 166 to cavity 180 and output duct 144 is a function onlysubstantially of the position of arm 234 relative to metering opening178 whenever the latter is the fuel flow controlling parameter. Asappropriate, a damping orifice 236 may be incorporated in pressuresensing line 182 to stabilize the movement of bypass valve poppet 168.

A uni-directional proportional solenoid 239 has an outer housing 238integral with plate 176 or otherwise affixed in stationary relationshipto housing 140. Disposed within the housing 238 is a coil 240, and acentrally arranged armature 242. Rigidly secured to form a portion ofarmature 242 is a central plunger shaft 244 which has an upper endengageable with lever arm 234. Linear gradient springs 246, 248 operablyextend between stops on housing 238 to engage associated shoulders onthe plunger shaft 244 to normally urge the latter to its de-energizedposition illustrated. Energization of the solenoid through appropriateelectrical lead lines 250 causes upward shifting of the armature 242 andplunger shaft 244 so that the latter engages and exerts an upward forceon lever arm 234 opposing and subtracting from the force exerted byspeeder spring 224 upon lever 226.

While the plunger shaft 244 could, if desired directly engage the leverarm 234; in the preferred form a "floating face" arrangement for arm 234is utilized. In this arrangement a floating flat poppet-type face 252 iscarried within arm 234 in alignment with metering opening 178. Thisfloating face is normally spring loaded toward the metering orifice, andthe upper end of plunger shaft 244 is engageable therewith. The purposeof floating face 252 is to compensate for manufacturing tolerances andto assure that a relatively flat surface is directly aligned withmetering opening 178 and lying perpendicular to the fluid flow therefromto assure proper metering of fuel thereacross. The spring 254 loadsfloating face 252 toward opening 178. Pivoting of arm 234 against spring254 to increase fuel flow is permitted until face 252 contacts the upperend of 245 of plunger 244. This stroking of arm 234 is quite limited butsufficient to create flow until face 252 contacts the upper end of 245of plunger 244. This stroking of arm 234 is quite limited but sufficientto create flow saturation of the annular orifice defined between opening178 and face 252.

Disposed on the opposite side of lever arm 234 from solenoid 239 is ahousing 256 of another directional, one-way solenoid 257 shown in FIGS.6B-6D. Solenoid 257 includes a coil 258, armature 260, and plunger shaft262 secured for movement therewith. Through appropriate stops, centeringsprings 264, 266 normally urge the plunger shaft 262 to the de-energizedposition illustrated. Upon energization of the coil 258 throughappropriate leader lines 268, the armature 260 and plunger shaft 262 areshifted downwardly such that the plunger shaft engages the lever arm 234in a manner exerting a force thereon tending to add to the force createdby speeder spring 224 and rotating lever 226 to shift arm 234 away fromopening 178. Housing 256 of solenoid 257 is rigidly secured such as bybolts 272 to securement plate 176. Similar to floating face 252, in thepreferred form the plunger 262 does not directly engage the lever arm234, but rather acts through a floating-type pin 272 to exert a force onarm 234. The pin 272 is pre-loaded by a spring 274 to give a floatingaction thereto in order to assure that plunger 262 can properly engageand exert a force on lever arm 234 regardless of variations inmanufacturing tolerances, and/or the position of lever 226 relative toits pivotal shaft 228.

Both solenoids are urged to their de-energized position by lineargradient springs, and unlike on-off, digital-type solenoids, variationin current and/or voltage input to their coils will cause an analogincremental positioning of the plunger 244 of solenoid 239, and willmove plunger 262 to either its FIG. 6-C or 6-D position.

The plunger 262 of solenoid 257 can be shifted away from itsde-energized FIG. 6-B state, to two different energized states shown inFIGS. 6-C and 6-D. One electrical input signal of preselected,intermediate power causes the armature 262 to shift to the FIG. 6-Cposition, moving plunger 262 until the face of its adjustable stop nut263 contacts the spring stop 267. This travel of plunger piston 262depresses plunger 272 and compresses spring 274 to shift arm 234 awayfrom opening 178 and increase fuel flow until gas generator speedincreases to a level corresponding to the signal force generated bysolenoid 257. Thus the plunger 272, spring 274 configuration assists inpermitting a less-than-maximum power signal to produce a force ofpreselected magnitude on arm 234.

Another electrical input signal of greater power causes the armature toshift to the end of its stroke with face 261 thereof contact theadjacent stop face 259 of the housing 256 as shown in FIG. 6-D. Thistravel causes piston plunger 262 to compress centering spring 266 andcause its lower end to come into direct contact with arm 234 and urgethe latter to permit maximum flow through the orifice presented betweenopening 178 and piston 252. As described in greater detail below,energization of solenoid 257 to its FIG. 6-D position is essentially afalse throttle signal duplicating the speed desired from the gasgenerator when the throttle is depressed to its maximum fuel flow,maximum power position.

Scheduling Valve 62

Referring now more particularly to FIGS. 7-11, scheduling valve 62generally includes a housing 276 which may be integral with bothhousings 140 and the stationary engine housing 110. Preferably housing276 is disposed in close proximity to both the fuel governor 60 and thecombustor 98. Housing 276 includes an internal bore 278 into which openthe two fuel ducts 142, 144 as well as the fuel line 64 and a lowpressure return conduit 280 which returns fuel back to the source.Mounted for longitudinal sliding and rotation within bore 278 is ametering valve 282 having "windowed" irregularly shaped openings 284,286 that open into the hollowed interior cavity 288 of valve 282. Fuelline 144 continuously communicates with interior cavity 288. Valve 282further includes an opening 290 in continuous communication with fuelline 64. Deceleration window 286 is in general alignment with fuel duct142, and acceleration window generally aligns with opening 290. Theparticular configuration of each of the windows 284, 286 is clearlyillustrated in FIGS. 10 and 11.

Metering valve 282 is urged in one longitudinal direction by a biasingspring 292 which reacts against the housing 276 through a spring stop294 acting on an alignment point 296 of a sealed block 298 mounted tohousing 276 such as by snap ring 300. The preferred construction asillustrated in FIG. 9; however, the alignment point arrangementpermitting rotation of valve 282 relative to housing 276 at the end ofspring 292 may alternately be accomplished via a ball 302 configurationas shown schematically in FIG. 7. At the opposite end of valve 282 is aspherical ball 304 permitting rotation of valve 282 relative to a piston306 carried in bore 278. Attached to housing 276 is a temperaturesensitive element 312, 308, for example a thermally responsive cylinder,whose longitudinal length varies with respect to the temperature imposedthereon by the gas or other fluid in the temperature sensing chamber 310within cylinder 312. The housing 276 is mounted relative to the enginesuch that a portion thereof, particularly cylinder 312 and theassociated chamber 310 are in communication with and maintained at thesame temperature, T₃.5, as the compressed air flow being delivered intothe combustor. Thermally insulative material 311 is incorporated asnecessary to avoid overheating of valve 62. For example the rightwardend of FIG. 9 and the perforated cylindrical wall 312 may be disposed atthe air inlet to the combustor and/or at the duct 96 carrying air fromthe recuperator 56 to combustor 98. In any case the scheduling valve isso arranged that cylinder 312 expands and contracts longitudinally withrespect to increase and decrease of combustor inlet temperature. Valve288 is operably engaged by the thermally responsive element 312 througha relatively non-thermally responsive ceramic rod 308. Accordingly,valve 288 is shifted longitudinally relative to input port 142 andopening 290 in relation to the sensed combustor inlet temperature. Thusthe metering fuel flow accomplised by window 284 is varied in relationto the sensed combustor inlet temperature as this window moveslongitudinally relative to opening 290.

Housing 276 further includes another transverse bore 314 which crossesand intersects generally with the longitudinal bore 276. Mounted forlongitudinal reciprocation within this transverse bore 314 is a rod andpiston configuration 316 which includes a pair of diaphragm-type seals318, 320 having outer ends rigidly secured to housing 276 by beingcompressed between the housing, an intermediate section 322 and aclosing plug 324 threadably or otherwise secured to housing 276. Theinner ends of the seals 320 are secured on the movable piston, rodconfiguration 316. The seal 320 in conjunction with the end closing plug324 define an interior pressure sensing chamber 326 to which one end ofthe piston 316 is exposed. Through a sensing line 328 the combustorpressure P₃.5 such as combustor inlet pressure is transmitted intochamber 326 to act upon one end of piston 316. At the opposite end ofbore 314, a helical coil biasing spring means 330, grounded to housing276 through a stationary stop 332, acts to urge the piston, rodconfiguration 316 in opposition to the pressure in chamber 326. Theopposite end 334 of the piston configuration 316 is vented toatmospheric pressure through an appropriate port 336. A sealschematically shown at 335, which may be of a structure like seals 318,320 and section 348, is also included at this opposite end 334. Thusgauge pressure in the combustor, i.e. the difference between ambientpressure and the absolute pressure maintained in combustor 98, acts uponpiston 316 to shift the latter within bore 314.

An arm 338 is threadably secured within a transverse bore in meteringvalve 282 at one end, and at its other end the rod 338 has a sphericalball 340 mounted thereon which is received in a groove 342 in rod,piston 316. It will therefore be apparent that shifting of piston, rod316 within bore 314 is translated into rotation of metering valve 282about its major longitudinal axis. Accordingly, the respective openingsbetween windows 284, 286 and the input ports 142 and opening 290 arealso varied in relation to the magnitude of gauge pressure in compressor98 by virtue of this rotational translation of metering valve 282.Groove 342 permits axial translation of arm 338 along with valve 282.While the rod, piston configuration 316 may be of varied arrangements,the preferred form as illustrated in FIG. 8 incorporates a threaded endsection 344 which acts through appropriate spaces 346 to compress andsecure the inner ends of seals 318, 320 to rod 316 through anintermediate section 348.

Thus, the scheduling valve acts as a mechanical analog computer inmultiplying the parameters of combustor pressure, P₃.5 and combustorinlet temperature, T₃.5, such that the positioning of valve 282 and thewindows 284, 286 is a function of the product quantity of combustorpressure multiplied by combustor inlet temperature.

Conventionally, as shown in FIG. 4 the controls for engine 30 furtherincludes a normally open, solenoid operated fuel sequencing solenoidvalve 350 as well as a manually or electrical solenoid operated shut-offvalve 352. These valves are disposed downstream of scheduling valve 62and in the preferred form may be included within and/or adjacent to thehousing 276 of scheduling valve 62.

The configuration of each of the windows 284, 286 as illustrated inFIGS. 8 and 9 are determined to solve a qualitative empirical formula ofthe following form:

    W.sub.f =(K.sub.1 -K.sub.2 T.sub.3.5) P.sub.3.5 +K.sub.3 T.sub.3.5

where: K₁, K₂ and K₃ are constants determined by the operationalcharacteristics of a particular gas turbine engine and are reflected bythe configuration of window 284 and associated opening 290.

By proper formulation of the window 284 and opening 290, the solution tothis equation as accomplished by scheduling valve 62 holds a constantmaximum turbine inlet temperature T₄ during all or at least a portion ofgas generator acceleration. Accordingly, when window 284 is thecontrolling parameter for fuel flow, scheduling valve 62 empirically bymechanical analog, controls fuel flow to maintain a substantiallyconstant turbine inlet temperature, T₄. Window 284 is the primaryoperating parameter during acceleration of the engine as described ingreater detail below. In contrast, window 286 is the controllingparameter during engine deceleration. While acceleration window 284 iscontoured to maintain a substantially constant maximum gas generatorturbine inlet temperature to provide maximum acceleration performancewithin the temperature limitations of the engine, the decelerationwindow 286 is contoured to limit and control fuel flow to prevent lossof combustion while affording substantial deceleration of the engine. Anextensive discussion of operation of a similar type of turbine inlettemperature computing valve, but which utilizes absolute rather thangauge combustor pressure, may be found in U.S. Patent Application No.689,339 of Rheinhold Werner, filed May 24, 1976, now U.S. Pat. No.4,057,960.

Vane Actuator 66

Details of the vane actuator control 66 are illustrated in FIGS. 12 and13. The vane control is hydromechanical in nature and generally includesa housing 354 having a pair of hydraulic pressure fluid supply ports356, 358 respectively receiving pressurized fluid from a high pressurepump source 360 and lower pressure pump source 362 each of which aredriven through the auxiliary power system of the engine. It isunderstood that the pumps 360, 362 may provide various other functionswithin the engines also such as lubrication.

Housing 354 has an internal, fluid receiving cylinder 364 in which isreciprocally mounted a piston 366 dividing the cylinder into opposedfluid pressure chambers. Rod or shaft 368 carried with piston 366extends exteriorly of housing 354 and is operably connected with thebell crank 130 of FIG. 13 so that, as described previously, linearreciprocation of rod 368 causes rotation of bell crank 130, ring gears126, 128 and the sets of variable guide vanes 120, 122.

High pressure hydraulic fluid from inlet port 356 is delivered into abore 370 within housing 354 located adjacent cylinder 364. Alsointersecting at spaced locations along bore 370 are high pressure fluidexhaust duct 372, and a pair of fluid work conduits 374, 376respectively communicating with the cylinder 364 on opposed sides ofpiston 366. Mounted for reciprocation within bore 370 is a directionalfluid control valve element 380 which is nominally positionable in theopen center position illustrated wherein high pressure hydraulic fluidfrom duct 356 communicates only with the exhaust port 372. A series ofcentering springs 382, 383, 384, 385 normally urge valve 380 to theposition shown. Valve 380 is of the four-way type and is shiftable onedirection to direct high pressure fluid from port 356 to conduit 374 andthe upper side of piston 366, while through conduit 376 the lower sideof the cylinder carrying piston 366 is vented to a low pressure return386 via bore 370, and communicating conduit 388. Valve 380 is shiftablein an opposite direction to direct pressure fluid from inlet 356 toconduit 376 and the lower side of piston 366, while conduit 374communicates with return 386 through a chamber 378 and return line 379.It will be noted that piston 366 cooperates with housing 354, such aswith a circular wall protrusion 390 thereof to prevent fluidcommunication between chamber 378 and cylinder 364.

Spring 382 acts to sense the position of piston 366 and the guide vaneangle, and as a feedback device in acting upon valve 380. The relativecompression rates of spring 382 in comparison to the springs 383-385provides a high gain response requiring large movement of piston 366(e.g. 14 times) to counteract as initial movement of valve 380 andreturn the valve to its center position. Thus it will be apparent thatpiston 366 acts in servo-type following movement to the movement of an"input piston" in the form of valve 380.

In bore 370 is a stepped diameter piston mechanism 392 shiftable inresponse to the magnitude of fluid pressure from a conduit 394 actingupon a shoulder 393 of piston 392. Piston 392 presents an adjustablestop for varying the compressive force of spring 383. Pressure acting onshoulder 393 is opposed by a spring 385. Slidably extending through thecenter of element 392 is a rod 395 which acts as a variably positionablestop upon the spring 384 extending between the upper end of rod 395 andvalve 380. Rod 395 is longitudinally shiftable in response to rotationof a fulcrum type lever 396 pivotally mounted to housing 354 at pivot398.

Vane actuator control 66 further includes another bore 400 in which ismounted a control pressure throttling valve 402. An input from thethrottle lever 184 of the engine acts to depress a variably positionablespring stop 404 to increase the force exerted by compression spring 406in urging valve 402 downwardly. Opposing spring 406 is a gradientcompression, helical coil spring 408. Valve 402 is variably positionableto meter hydraulic flow from port 358 to conduit 410. It will be notedthat conduit 410 also communicates with the lower end of throttlingvalve 402 via a conduit 412 having a damping orifice 414 therein.Conduit 410 leads to the larger face of a stepped piston 416reciprocally mounted within another bore 418 in housing 354. One end onbore 418 is in restricted fluid communication with return 387 through anorifice 419. The smaller diameter section of stepped piston 416 receivespressurized fluid from conduit 420. Through an appropriate exhaustconduit 424 the intermediate section of the stepped piston, as well asthe upper end of valve 402 are exhausted to low pressure return 386through the conduit 388.

Conduit 420 provides a hydraulic signal indicative of the speed of thepower turbine shaft 82. In this connection, the vane actuator includes anon-positive displacement type hydraulic pump, such as a centrifugalpump 422 mounted to and rotated by power turbine shaft 82. Being anon-positive displacement type pump, the pump 422 delivers pressurizedhydraulic flow through conduit 420 such that the pressure maintained onthe smaller diameter of stepped piston 416 is a square function of thespeed of power turbine shaft 82. Similarly, the action of throttlingvalve 402 develops a pressure on the large diameter of piston 416 inrelation to a desired or selected speed reflected by the position of thethrottle 184.

The valve 402 and piston 416 act as input signal means and as acomparator to vary the compressive force of spring 384 as a function ofthe difference or error between actual power turbine speed and the powerturbine speed requested by throttle position. The requested N_(pt) isgraphically illustrated in FIG. 19.

The vane actuator control 66 further includes a linear, proportionalsolenoid actuator 426 operably connected by electrical connector lines427 to electronic control module 68. Actuator 426 includes a housing 428enclosing a coil 430, and a centrally arranged armature which carriestherewith a hydraulic directional control valve 432. Valve 432 isnormally urged upwardly by spring 434 to the position communicatingconduit 394 with return 386. Valve 432 is proportionally shiftabledownwardly in response to the magnitude of the energization signal toproportionally increase communication between conduits 372 and 394 whiledecreasing communication between conduit 394 and drain. As a result,pressure in conduit 394 increases proportionately to the magnitude ofthe electronic signal, such pressure being essentially zero in theabsence of an energization signal to solenoid 426. It will be noted thatminimum pressure in conduit 394 allows springs 383 and 385 to exertmaximum upward force on valve 380, and that increasing pressure inconduit 394 shifts piston 392 downwardly to reduce the force exerted bysprings 383, 385 upon valve 380, thus developing an override force inthe form of reduced force from spring 383.

In the absence of an electrical signal to solenoid 426 minimum pressureis exerted on shoulder 393 causing the guide vanes to be controlled bypower turbine speed. Thus, the guide vanes during start-up are at theirFIG. 14 position and at other conditions of engine operation arenormally urged to maximum power, FIG. 15 position.

As shown in FIG. 18, vane actuator 66 is operable to vary guide vaneangle, B, from 0° to +20° to alter the positive incidence of gas flowupon the power turbine blades and thus alter power transmitted from thegas flow to rotate the power turbine wheels in a direction transmittingmotive power to the vehicle. The vane actuator 66 is also operable toshift the guide vanes to a negative incidence position and modulate theguide vane position within zone "d" of FIG. 18. In these negativeincidence positions, gas flow is directed to oppose and thus tend todecelerate the rotation of the power turbine wheels.

Electronic Control 68

A portion of the control logic of the electronic control module 68 isillustrated in FIG. 17. The electronic control module receives inputelectrical signals indicative of power turbine speed (N_(pt)) through achopper 436 secured to power turbine shaft 82 and an appropriatemagnetic monopole 438 which transmits an electronic signal indicative ofpower turbine speed through lead line 440. Similarly, gas generatorspeed, N_(gg), is sensed through a chopper 442, monopole 444 and leadlines 446. Transducers 448, 450, and 452 respectively generateelectrical input signals indicative of the respective temperature sensedthereby, i.e. compressor inlet temperature T₂, turbine inlet temperatureT₄, and turbine exhaust temperature T₆. As illustrated these temperaturesignals are transmitted through lines 454, 456 and 458. The electroniccontrol module also receives from an ambient pressure sensor 460 andassociated line 462 an electrical signal indicative of ambient pressureP₂. The electronic control module further receives from an appropriatesensing device an electrical signal through lines 464 indicative ofthrottle 184 position, "a." Also, a switch 466 is manually settable bythe vehicle operator when power feedback braking (described more ingreater detail below) is desired. A transducer 544 generates a signal toan inverter 546 whenever the variable guide vanes are moved past apredetermined position B*.

The electronic control module includes several output signals toenergize and/or de-energize the various logic solenoids and relaysincluding solenoid 518 through line 519, solenoid 257 through line 268,fuel sequencing solenoid 350 through associated line 351, fuel trimsolenoid 239 through line 250, and the vane solenoid 426 through line427. The electronic control module includes function generators 514, 550and 552. Box 514 is denoted as a "flat rating and torque limiting"function and generates a signal indicative of maximum allowable gasgenerator speed as a function of ambient conditions T₂ and P₂ and powerturbine speed N_(pt). Element 550 transforms the throttle positionsignal "a" into an electronic gas generator speed request signal, andfunction generator 552 produces a signal as a function of gas generatorspeed N_(gg) from line 446. The module further includes comparators 497,534, 540, 554, 556 as well as the logical elements 498, 500 and 538. Thelogical elements are of the "lowest wins" type, i.e. they pass thealgebraically lowest input signal.

The logic element 498 selects from the signals 536 and 542 which havebeen generated in comparators 534 and 540 indicating the amount of overor undertemperature for T₄ and T₆. An additional input from 456 isprovided to logic element 498 so as to provide an indication ofexcessive T₄ figures in the case of a failed T₄ sensor signal. The logicelement 500 receives inputs from 497 and 498. Comparator 497 comparesthe electronic speed request with the actual gas generator speed 446 todetermine if the engine has been requested to accelerate or is in steadystate. The output of logic element 500 is fed to inverter 546,generating an appropriate signal in solenoid driver 558 which then movestrim solenoid 426 a distance proportional to the magnitude of signal427.

The logic element 538 receives its inputs from comparators 554 and 556,logic element 498 and a differentiator 548. As noted, logical element498 indicates the lower of the two temperature errors T₄ and T₆. Theoutput of comparator 556 is the error between the operator requestedpower turbine speed N_(pt) and the actual power turbine speed N_(pt).The output of comparator 554 is indicative of the difference between themaximum allowable gas generator speed determined by function generator514 and the actual gas generator speed 446. The logic element 538selects the algebraically lowest signal and outputs it to solenoiddriver 560 with an output on line 250 which is passed on to the governorreset decrease solenoid 239 in the fuel control 60.

As depicted in FIG. 17, the electronic control module includes acomparator 468 and synthesizers or function generators 470, 472 and 474.Function generator 470 produces an output signal in line 478 indicativeof whether the difference between power turbine speed and gas generatorspeed is less than a preselected maximum such as five percent. Functiongenerator 472 produces a signal in line 480 showing whether or not powerturbine speed is greater than gas generator speed, while functiongenerator 474 generates a signal in lines 482 showing whether or not gasgenerator speed is greater than 45 percent of its maximum speed. Thecontrol logic further includes function generator 486 and 488 whichrespectively generate signals in associated line 490 and 492 showingwhether or not transmission input speed is above a preselected minimum"e" and whether throttle position is below a preselected throttleposition a*. Throttle position "a" is obtained from a suitable positionsensor such as a variable resistance potentiometer. Thus, output signal464 is indicative of throttle position "a."

The electronic control module further includes the logical gates 502,504, 506, 508 and 562. Logical AND gate 502 receives inputs from line478 and AND gate 506 to produce an output signal to solenoid driver 516to activate power feedback clutch 84. Logical AND gate 506 receives itsinputs from line 482, switch 466 and line 492 and produces an inputsignal to AND gates 502 and 504. Logical AND gate 504 receives an inputfrom line 480 and the inverted input from line 478. Its output generatesa 50% gas generator speed signal and also enables solenoid driver 564through OR gate 562 to produce the "a" signal in line 268 which is theresult of a constant 50% signal plus the output of element 566. Signal268 then activates the governor reset increase solenoid 257 in the fuelcontrol 60. Logical AND gate 508 receives its inputs from lines 490 and492. Its output signal generates a 20% gas generator signal throughfunction generator 568 which, added to the constant 50% signal by summer570 results in a fast idle signal (70% gas generator speed) to thegovernor reset increase solenoid 257. The output of AND gate 508 alsogenerates the enable signal to solenoid driver 564.

Power Feedback Clutch 84

While various forms clutches could be utilized for power feedback clutch84, the preferred form shown in FIG. 3 comprises a "wet" typehydraulically actuated clutch which includes a shaft 520 from the geartrain 78 associated with gas generator shaft 76, and a shaft 522interconnected with the gear train 80 associated with the power turbineoutput shaft 82. The clutch operates in a continual bath of lubricatingcooling fluid. The gas generator shaft 520 drives a plurality of discs524, which are interposed in discs 526 connected to the output shaft522. The clutch actuator is in a form of a solenoided operateddirectional hydraulic control valve 518 which, in the energized positionillustrated, ports pressurized fluid such as from source 362 into afluid pressure chamber 528 to urge piston 530 against the urgings of areturn spring 532 to force the plates 524, 526 into interengagement suchthat the power from shaft 522 may be fed back to a gas generator shaft520 to assist in braking. When the solenoid actuator 518 isde-energized, the chamber 528 is exhausted to a low pressure drain topermit the spring 532 to shift piston 530 away from the position shownand disengage the plates 524, 526.

OPERATION Starting

In a conventional manner start motor 72 is electrically energized toinitiate rotation of gas generator drive shaft 76 and the input shaft152 of fuel governor 60. The control module 68 energizes the normallyopen fuel sequence solenoid 350, and solenoid 352 is also in an openposition to clear fuel line 64 for delivery to the combustor. Asnecessary, an assist pneumatic pump 74 delivers pressurized air intocombustor 98 along with the action of ignition plugs 100. Motor 72 isutilized to drive the various components described until the gasgenerator section reaches its self-sustaining speed, normally in therange of approximately 40% of maximum rated gas generator speed.

During initial rotation and starting of the engine, the low speed ofrotation of fuel governor drive shaft 152 cannot overcome the bias ofspeeder spring 224, and thus fuel lever 226 is disposed away from andclearing orifice 178 to permit fuel flow from line 166 to output line144. Also during this initial starting, the combustor temperature (T₃.5)and combustor pressure (P₃.5) are both relatively low such thatscheduling valve 62 also permits significant fuel flow through line 64to the combustor.

Low Idle

As gas generator shaft 76 speed climbs beyond the self-sustaining speed,start motor 72 is shut off and the combustion process permitsself-sustaining operation of the gas generator. Speeder spring 224 isnormally set to maintain a low idle value of approximately 50% ofmaximum gas generator rated speed. Accordingly, the mechanical flyweightgovernor operates in opposition to speeder spring 224 to adjust fuellever 226 and maintain fuel flow through orifice 178 to hold gasgenerator speed at a nominal 50% of maximum. This 50% low idle speed iseffective whenever proportional solenoid 257 is in the de-energizedstate illustrated in FIG. 6.

The electronic control module 68 normally maintains solenoid 257 in thede-energized state to select the low idle gas generator speed wheneverthe transmission input shaft speed of shaft 36, as sensed by speedsensor 48, is rotating. Such normally occurs whenever the clutch 34 isengaged with transmission 38 in its neutral position, or whenever thevehicle is moving regardless of whether or not the clutch 34 is engagedor disengaged. Accordingly, during idling when not anticipatingacceleration of the engine, the comparator 486 of the electronic controlmodule 68 notes that the speed of shaft 36 is above a pre-determinedminimum, "e", such that no signal is transmitted from comparator 486 toAND gate 508. Solenoid 257 remains de-energized, and the gas generatorspeed is controlled by the governor to approximately 50% its maximumspeed.

High Idle

Maximum power is normally required to be developed from an enginedriving a ground vehicle upon initiating acceleration of the vehiclefrom a stationary or substantially stationary start. As a naturalconsequence of normal engine operator action upon starting from astationary start, transmission input shaft 36 comes to a zero or verylow rotational speed as clutch 34 is disengaged while gear shift lever46 is articulated to shift the transmission into gear. Once the speed ofshaft 36 drops below a predetermined speed, "e", comparator 486 of theelectronic control module generates an output signal to AND gate 508.Since accelerator lever 184 is still at its idle position, the sensorassociated with line 464 generates a signal to energize comparator 488and also send a positive signal to AND gate 508. The output of AND gate508 energizes function generator 568 to add 20% to the constant idlecommand of 50%, so that summer 570 provides a 70% command signal tosolenoid driver 564 that has been abled through the output of AND gate508 and OR gate 562. Accordingly, solenoid 257 is energized by anappropriate current signal through line 268 to shift to its FIG. 6Cposition. In this position the solenoid 257 has been sufficientlyenergized to drive shaft 262 and plunger 272 downwardly as viewed inFIG. 6C and exert a force on fuel lever 226 tending to rotate the latteraway from and increase the size of orifice 178. The additional forceexerted by solenoid 257 is sufficient to increase fuel flow throughorifice 178 to increase gas generator speed to a predetermined higherlevel, such as 70% of maximum gas generator speed. The flyweightgovernor operates to hold the gas generator speed constant at thislevel.

In this manner, a the idle speed of the gas generator section is resetto a higher value in anticipation of a required acceleration such thatmore power will be instantly available for accelerating the vehicle. Atthe same time, when acceleration is not anticipated, as determined bywhether or not transmission input shaft 36 is rotating or stationary,the electronic control module 68 is operable to de-energize solenoid 257and reduce gas generator speed to a lower idle value just above thatnecessary to maintain a self-sustaining operation of the gas generatorsection. In this manner power necessary for acceleration is availablewhen needed; however during other idling operations the fuel flow andthus fuel consumption of the engine is maintained at a substantiallylower value. This is accomplished by producing a signal, minimum speedof shaft 36, which is anticipatory of a later signal (rotation ofaccelerator lever 184) requesting significant increase in powertransmitted to drive the vehicle.

Acceleration

Acceleration of the gas turbine engine is manually selected bydepressing the accelerator 184. To fuel governor 60 this generates a gasgenerator section speed error signal in that the depression of lever 184rotates shaft 192 to increase compression of speeder spring 224 beyondthat force being generated by the mechanical flyweight speed sensor.Fuel lever 226 rotates in a direction substantially clearing the opening178 to increase fuel flow to the combustor.

At the same time, depression of throttle lever 184 generates a powerturbine section speed error signal to vane actuator control 66. Moreparticularly, depression of throttle 184 compresses spring 406 to shiftvalve 402 downwardly and increase the pressure maintained in chamber 418substantially beyond that being generated by the hydraulic speed signalgenerator of pressure developed by pump 422 and exerted on the otherside of the step piston 416. Accordingly, lever 396 is rotated generallyclockwise about its pivot 398 in FIG. 12, allowing downward retraction,if necessary, of plunger 395 and reduction of compression on spring 384.

Summer 497 of the electronic control module determines a large disparitybetween accelerator position and gas generator speed to develop anelectronic signal to element 500 overriding other signals thereto andreducing the signal in line 427 to zero to de-energize the solenoid 426of guide vane control 66. The spring bias urges plunger 430 and valve432 to the position shown in FIG. 12 to minimize hydraulic pressuredeveloped in conduit 394 and exerted on piston shoulder 393. Asdiscussed above in the vane control 66 description, springs 382-385position valve 380 to cause following movement of piston 366to itsnominal or "neutral" position. In this position vane piston 366 and rod368, the guide vanes 120 are disposed in their FIG. 14 position whereinthe gas flow from the combustor is directed onto the power turbine vanesin a manner minimizing power transfer to the power turbine vanes. Moreparticularly, the guide vanes 120 are disposed in their FIG. 14 positionto reduce the pressure drop or pressure ratio across turbine blades 117to a minimum value, this position corresponding to the 0° position ofFIG. 18.

Since the nozzles 104 maintain the combustor 98 in a choked condition,this reduction in pressure ratio across the turbine blades 117 creates asubstantial increase in pressure ratio across the radial inflow turbine102 of the gas generator section. Accordingly positioning of the guidevanes in their FIG. 14 position by allowing the springs 382-385 toposition valve 380 and piston 366 in its "neutral" position, alters thepower split between the gas generator turbine 102 and the power turbines116, 118 such that a preselected maximum portion of power from themotive gas flow is transmitted to the gas generator turbine 102. As aresult, maximum acceleration of the gas generator section from eitherits low or high idle setting toward its maximum speed is achieved. Asnoted previously, the requirement for impending acceleration has beensensed, and the engine is normally already at its high idle setting sothat gas generator speed promptly nears its maximum value.

As gas generator speed increases, the combustor pressure P₃.5accordingly increases. This causes rotation of the metering valve 282 ofthe fuel schedule control 62 to increase the amount of overlap betweenacceleration schedule window 284 and opening 298 in the fuel schedulingvalve. Increase in this opening causes a consequent increase in fuelflow to combustor 98 and an ultimate resulting increase in the inlettemperature T₃.5 through the actions of recuperator 56.

To the operation of engine 30, increase in T₃.5 is in practical effectthe same as a further fuel flow increase. Accordingly, in solving theabove described equation the window 284 shifts to reduce fuel flow withincreasing T₃.5 to produce an "effective" fuel flow, i.e., one combiningthe effects of actual fuel flow and inlet temperature T₃.5, at thesensed gauge pressure P₃.5 to produce a desired combustor exhaust or gasgenerator turbine inlet temperature T₄.

This increase in fuel flow created by the rotation of valve 282 and ascompensated by axial translation of the valve provides an "effective"fuel flow that increases power developed and transmitted from the gasflow to gas generator turbine 102. This then causes another increase ingas generator speed, and combustor pressure P₃.5 again increases.Scheduling valve thus acts in regenerative fashion to further acceleratethe gas generator section. As noted previously, the scheduling valve isso contoured to satisfy the equation discussed previously and allowcontinued increase in P₃.5 while maintaining combustor outlettemperature T₄ at a relatively constant, high value. In this manner thegas generator section is accelerated most rapidly and at maximumefficiency since the turbine inlet temperature T₄ is maintained at ahigh, constant value.

While the acceleration window 284 and opening 290 may be relativelyarranged and configured to maintain a constant T₄ throughoutacceleration, a preferred form contemplates maintaining a substantiallyconstant T₄ once the power turbine has initiated rotation, whilelimiting turbine outlet or recuperator inlet temperature during a firstpart of the accleration operation. In this manner excessive T₆ isavoided when the power turbine section is at or near stall. Morespecifically, it will be noted that upon starting acceleration of thevehicle, the free power turbine section 54 and its shaft 82 arestationary or rotating at a very low speed due to the inertia of thevehicle. Thus there is little temperature drop in the gas flow whileflowing through the power turbine section, and the recuperator inlettemperature T₆ starts approaching the temperature of gas flow exitingthe gas generator radial turbine 102. If combustor exhaust or gasgenerator turbine inlet temperature T₄ is maintained at its maximumconstant value at this time, it is possible that T₆ may becomeexcessively high in instances of high inertial load which lengthens thetime of this substantial "stall" condition on the power turbine section.Of course, as the power turbine section overcomes the inertia andreaches higher speeds, temperature drop across the power turbinesincreases to hold down recuperator inlet temperature T₆.

For such free turbine type engines, relatively complicated and expensecontrols, electronic and/or mechanical, are normally expected in orderto avoid excessive T₆ while providing responsive acceleration under theconditions in question. An important discovery of the present invention,as embodied in scheduling valve 62, is in providing an extremely simple,economical, mechanical structure capable of limiting T₆ during thecritical turbine section stall period but yet still promoting veryresponsive engine acceleration. At the same time this improvedarrangement has eliminated the need for compensation for substantialvariations in ambient pressure and thus the need to compensate for thevariations in altitude that would be expected to be encountered by aground vehicle. In this connection it would be expected that absolutecombustor pressure P₃.5 must be the parameter in solving the equationdescribed previously such that the scheduling valve could reliablycompute the turbine inlet temperature T₄ created by a particularcombination of combustor pressure, P₃.5, and inlet temperature, T₃.5.

However, a discovery of the present invention is that by properselection of the constants K₁, K₂ as embodied in the size andconfiguration of openings 284, 290, and by utilization of combustorgauge pressure rather than combustor absolute pressure, mechanicallysimple and economical structure with minimum control complexity canaccomplish the desired control of both T₆ and T₄ during acceleration.Window 284 and opening 290 are relatively arranged such that when valve282 rotates to a minimum P₃.5, a slight overlap remains between thewindow and opening. Thus, a minimum fuel flow, W_(f), is maintained atthis condition which is a function of T₃.5 since valve 282 is stillcapable of translating axially. This gives rise to the third term, K₃T₃.5, in the equation set forth above and dictates an initial conditionof fuel flow when window 284 becomes the controlling fuel flow parameterupon starting acceleration.

The constants K₁, K₂ are chosen, their actual values being determined bythe aerodynamic and thermodynamic characteristics of the engine, suchthat at a preselected value, P₃.5 *, intermediate the maximum andminimum values thereof, the acceleration window controls fuel flow tomaintain a constant T₄. At combustor pressures below this preselectedvalue, the acceleration window provides fuel flow allowing T₄ to reducebelow the preselected maximum desired level therefor. It has been foundthat an inherent function of using gauge combustor pressure rather thanabsolute pressure, in combination with these chosen values of K₁, K₂ anda preselected minimum fuel flow at minimum P₃.5 as determined by K₃, isthat fuel flow is controlled by the acceleration window to preventrecuperator inlet temperature T₆ from exceeding a preselected value.This approach still utilizes the simple geometry of window 284 and 290,both rectangles, that mechanicaly compute the product of T₃.5 multipliedby P₃.5. Accordingly, at pressures lower than P₃.5 * which arecharacteristic of the conditions under which the turbine section"stalling" occurs, the utilization of gauge combustor pressure preventspotentially damaging excessive T₆. The design point for window 284 is,of course, the condition of maximum vehicle inertia experienced onturbine shaft 82, lesser values of such inertia naturally permittingmore rapid turbine shaft speed increase and less time in the "stalling"condition above described.

From inspection of the equation solved by valve 282 it will be apparentthat fuel flow W_(f) is a linear or straight line function of P₃.5 asshown in FIG. 20, with a slope determined by K₁ and K₂, an interceptspecified by K₃, and passing through the point producing the preselectedturbine inlet temperature T₄ at the selected intermediate value P₃.5 *.Of course, a family of such straight line curves of W_(f) vs. P₃.5results for different values of T₃.5 While, if desired, curve fitting ofwindow 284 and opening 290 could be utilized to maintain T₄ at preciselythe same value at pressures at and above the preselected intermediateP₃.5 *, in the preferred form compound curvature of the window andopening is not utilized. Instead, the window and opening are ofrectangular configuration thus permitting T₄ to increase very slightlyat combustor pressures above P₃.5 *. However, it has been found thatsuch arrangement affords an excellent, practical approximation to thetheoretically desired precisely constant T₄, resulting in practicaleffect in maintaining a substantially constant T₄ at a desired maximumvalue once combustor gauge pressure exceeds the preselected level P₃.5*. Accordingly, the present invention inherently limits recuperatortemperature T₆ to solve the problem of recuperator overheating whenstarting to accelerate a high inertial load, yet still maintains amaximum T₄ for high engine efficiency throughout the remainder ofacceleration once the inertia is substantially overcome and for themajority of time during acceleration. At the same time, and contrary towhat might normally be expected, it has been found that the need foraltitude compensation is obviated since there exists a minimum fuel flowat minimum combustor pressure, which minimum fuel flow varies linearlywith combustor inlet temperature T₃.5. Thus the present inventionprovides a simple mechanical solution to the interdependent and complexproblems of limiting two different temperatures T₄, T₆ for differentpurposes, i.e., avoiding recuperator overheating while affording highengine operating efficiency and thus high responsive acceleration.

As the gas generator continues to accelerate, the flyweight governor 208of the fuel governor 60 beings exerting greater downward force tocounteract the bias of speeder spring 224. Accordingly, the fuel lever226 begins rotating in a generally counter-clockwise direction in FIG. 6to begin metering fuel flow through opening 178. Once the opening 178becomes smaller than that afforded by metering window 284 in scheduldingvalve 62, the operation of the scheduling valve is overridden and thefuel governor 60 begins controlling fuel flow to the combusltor in amanner trimming gas generator speed to match the speed selected by therotation of the shaft 192 associated with the acceleration lever 184 inthe fuel governor 60.

Similarly, this increase in gas generator speed is sensed in theelectronic control module 68 by summer 497 such that once gas generatorspeed N_(gg) approaches that selected by the position of the acceleratorpedal as sensed electronically through line 464, the override signalgenerated by summer 497 is cut off. In response, element 500 is allowedto generate a signal energizing the proportional solenoid 426 of theguide vane control 66. Valve 432 associated with solenoid 426 is shiftedto increase pressure exerted upon piston shoulder 393 to permit thepiston 366 and the guide vanes to shift from the FIG. 14 dispositionthereof towards the FIG. 15 position. This shifting of the guide vanesfrom the FIG. 14 to the FIG. 15 position again alters the work splitbetween the gas generator turbine 102 and the power output turbines 116,118 such that greater power is developed across the output turbines andtransmitted to output shaft 82 while a lesser portion is transmitted toturbine 102.

Thus it will be apparent that acceleration of the engine and vehicleoccurs by first altering the work split so that maximum power isdeveloped across the gas generator turbine 102, then increasing fuelflow along a preselected schedule to regeneratively further increasepower developed across the gas generator while maintaining turbinecombustor exhaust temperature T₄ at a substantially constant,preselected maximum. Once substantial acceleration of the gas generatorsection has been accomplished, the guide vanes are then rotated to alterthe power or work split so as to develop a greater pressure ratio acrossand transmit more power to the power turbines 116, 118 and the poweroutput shaft 82.

Cruise

During normal cruise operation (i.e., traveling along at a relativelyconstant speed or power output level) the guide vane control 66 actsprimarily to alter the work split between the gas generator turbine 102and the power output turbines 116, 118 so as to maintain a substantiallyconstant combustor exhaust temperature T₄. This is accomplished by theelectronic control module which includes a summer 534 developing anoutput signal in line 536 to the logic box 498 indicative of thedifference between the actual and desired turbine inlet temperature T₄.More particularly, solenoid 426, as discussed previously, is maintainednormally energized to generate maximum pressure upon the piston shoulder393 of the guide vane actuator. For instance, assuming that T₄ is abovethe preselected desired value thereof, a signal is generated to line 536and element 498 to reduce the magnitude of the electric signal throughline 427 to solenoid 426. Accordingly, the spring bias 434 of thesolenoid begins urging valve 432 in a direction reducing fluidcommunication between conduits 372 and 394 while correspondinglyincreasing communication between conduit 394 and exhaust conduit 386.The reduction in pressure exerted upon piston 393 accordingly allowsspring 385 to increase the spring bias of spring 383 to cause upwardtravel of valve 380 and corresponding downward travel of piston 366 todrive the vanes backwards from their FIG. 13 disposition (+20° positionof FIG. 18) toward a wider open position increasing the area ratio andreducing the pressure ratio across the vanes of the turbines 116, 118.Accordingly, in response to T₄ over-temperature, the guide vanes areslightly opened up to reduce the pressure ratio across the turbines 116,118. In response the increased pressure ratio across gas generatorturbine 102 causes an increase in gas generator speed. Such increase ingas generator speed is then sensed by the flyweight governor 208 of thefuel governor 60 to cause counter-clockwise rotation of fuel lever 226and reduce fuel flow through opening 178. The reduction in fuel to thecombustor 98 accordingly reduces the combustor exhaust or turbine inlettemperature T₄ toward the preselected value thereof. Thus, the guidevane control operates to adjust the guide vanes as necessary and causesa consequent adjustment in fuel flow by the fuel governor 60 due tochange in gas generator speed N_(gg) so as to maintain the combustorexhaust temperature T₄ at the preselected, maximum value. It will beapparent also from the foregoing that reduction in turbine inlettemperature T₄ below the preselected desired value thereof causes acorresponding movement of the guide vanes 120, 122 to increase thepressure ratio across the power turbines 116, 118. Accordingly thiscauses a reduction in pressure ratio across gas generator turbine 102 toreduce gas generator speed. In response the fuel governor shifts fuellever 226 in a clockwise rotation as viewed in FIG. 6 to increase fuelflow to the combustor and thus increase turbine inlet temperature T₄back to the desired value. It will be apparent that the change in guidevane position also directly alters the combustor exhaust temperature T₄due to the difference in air flow therefrom; however, the majoralteration of combustor exhaust temperature is effected by altering thefuel flow thereto as described above.

During the cruise operation therefore, it should now be apparent thatfuel governor 60 acts to adjust fuel flow in such a manner as tomaintain a gas generator speed in relation to the position of theaccelerator lever 184. Clearly, the fuel governor 60 operates inconjunction with or independently of the vane control 66, dependent onlyupon the gas generator speed N_(gg).

While the electronic control module operates the guide vane controlsolenoid 426 to trim turbine inlet temperature T₄ during cruise, thehydromechanical portion of the guide vane control 66 acts in a moredirect feedback loop to trim the speed of power turbine output shaft 82.More particularly, the actual power turbine speed as sensed by thepressure developed in line 420 is continuously compared to theaccelerator lever position as reflected by the pressure developed inline 410. A graphical representation of the action of valve 402 andpiston 416 in compressing spring 384 and requesting different desiredpower turbine speeds N_(pt) in relation to the throttle position, a, isshown in FIG. 19. Thus, in response to an increase in speed of powerturbine shaft 82 beyond that selected by the rotation of acceleratorlever 184, pressure at the lower diameter of piston 416 becomessubstantially greater than that on the larger face thereof to rotatelever 396 so as to increase compression of the biasing spring 384 actingon valve 380. The resulting upward movement of valve 380 causes acorresponding downward movement of piston 366 and accordingly shifts theguide vanes toward the FIG. 14 position, i.e., opens the guide vanes toincrease the area ratio and reduce the pressure ratio across the vanes117, 119 of the two power turbine wheels. This reduces the powertransmitted from the gas flow to the power turbine wheel and thus causesa slight decrease in power turbine output shaft speed back to thatselected by the accelerator lever 184. It will be apparent that wheneverthe speed of the power turbine shaft 82 is below that selected byaccelerator lever 184, the compression of spring 384 is reduced to tendto increase the pressure ratio across the power turbine vanes 117, 119to tend to increase power turbine speed N_(pt).

The portion of vane control 66 for trimming power turbine speed inrelation to accelerator position is preferably primarily digital inaction since as shown in FIG. 19, a small change in throttle leverposition increases the requested N_(pt) from 25% to 100%. The actions ofvalve 402, piston 416 and plunger 395 are such that when the acceleratoris at a position greater than a*, this portion of the controlcontinually requests approximately 105% power turbine speed N_(pt).Through a small amount of rotation of the accelerator below a*, thecontrol provides a request of power turbine speed proportional to theaccelerator position. Positioning of the accelerator to an angle belowthis small arc causes the control to request only approximately 25% ofmaximum N_(pt).

Thus, in normal cruise the guide vanes control operates in conjunctionwith the fuel governor to maintain a substantially constant turbineexhaust temperature T₄ ; fuel governor 60 operates to trim gas generatorspeed N_(gg) to a value selected by the accelerator lever 184; and thehydromechanical portion of guide vane 66 operates to trim power turbineoutpt speed N_(pt) to a level in relation to the position of acceleratorpedal 184. It will further be apparent that during the cruise mode ofoperation, the orifice created at opening 178 of the fuel governor issubstantially smaller than the openings to fuel flow provided in thescheduling valve 62 so that the scheduling valve 62 normally does notenter into the control of the engine in this phase.

Safety Override

During the cruise or other operating modes of the engine discussedherein, several safety overrides are continually operable. For instancesolenoid 239 of the fuel governor 60 operates to essentially reduce orcounteract the effect of speeder spring 224 and cause a consequentreduction in fuel flow from orifice 178 by exerting a force on fuellever 226 tending to rotate the latter in a counter-clockwise directionin FIG. 6. As illustrated in FIG. 17, the electronic control moduleincludes a logic element 538 which is responsive to power turbine speedN_(pt), gas generator speed N_(gg), turbine inlet temperature T₄, andturbine exhaust or recuperator inlet temperature T₆. Thus if turbineinlet temperature T₄ exceeds the preselected maximum, a proportionalelectrical signal is transmitted to lines 250 to energize solenoid 239and reduce fuel flow to the engine. Similarly, excessive turbine exhausttemperature T₆ results in proportionately energizing the solenoid 239 toreduce fuel flow to the combustor and thus ultimately reduce turbineexhaust temperature T₆. Also, logic element 438 is responsive to powerturbine speed so as to proportionately energize solenoid 239 wheneverpower turbine speed exceeds a preselected maximum. Similarly, theelectronic control module operates to energize solenoid 239 whenever gasgenerator speed exceeds a preselected maximum established by functiongenerator 514 as a function of P₂, T₂ and N_(pt). Normally thepreselected maximum parameter values discussed with regard to thesesafety override operations, are slightly above the normal operatingvalues of the parameters so that the solenoid 239 is normally inoperableexcept in instances of one of these parameters substantially exceedingthe desired value thereof. Thus, for instance, during a cruise mode ofoperation or "coasting" when the vehicle is traveling downhill beingdriven to a certain extent by its own inertia, the solenoid 239 isoperable in response to increase of power turbine output shaft 82 beyondthat desired to cut back on fuel flow to the combustor to tend tocontrol the power turbine output speed.

While as discussed previously with regard to the cruise operation of thevehicle, the guide vane control normally is responsive to combustorexhaust temperature T₄ as reflected in the signal generator by element435, the logic element 498 is also responsive to the turbine exhausttemperature T₆ in comparison to a preselected maximum thereof asdetermined by summer 540 which generates a signal through line 542 toelement 498 whenever turbine exhaust temperature T₆ exceeds thepreselected maximum. Logic element 498 is responsive to signal fromeither line 542 or 536 to reduce the magnitude of the electronic signalsupplied through line 427 to solenoid 426 and thus reduce the pressureratio across the turbine wheels 116, 118. As discussed previously, thischange in pressure ratio tends to increase gas generator speed and inresponse the fuel governor 60 reduces fuel flow to the combustor so thatturbine exhaust temperature T₆ is prevented from increasing beyond apreselected maximum limit.

As desired, the solenoid 239 may be energized in response to otheroverride parameters. For instance, to protect the recuperator 56 fromexcessive thermal stresses, the logic element 538 may incorporate adifferentiator 548 associated with the signal from the turbine exhausttemperature T₆ so as to generate a signal indicative of rate of changeof turbine exhaust temperature T₆. Logic element 538 can thus generate asignal energizing solenoid 239 whenever the rate of change of turbineexhaust temperature T₆ exceeds a preselected maximum. In this mannersolenoid 239 can control maximum rate of change of temperature in therecuperator and thus the thermal stress imposed thereon. Similarly, thelogic element 538 may operate to limit maximum horsepower developedacross the power turbine and/or gas generator shafts.

Gear Shift

Because turbine engine 30 is of the free turbine type with a poweroutput shaft 82 not physically connected to the gas generator shaft 76,the power turbine shaft 82 would normally tend to greatly overspeedduring a gear shifting operation wherein upon disengagement of the driveclutch 34 to permit gear shifting in box 38, substantially all inertialretarding loads are removed from the power turbine drive shaft 82 andassociated power shaft 32. Of course, during normal manual operationupon gear shifting, the accelerator lever 184 is released so that thefuel governor 60 immediately begins substantially reducing fuel flow tocombustor 98. Yet because of the high rotational inertia of the powerturbine shaft 82 as well as the high volumetric gas flow thereacrossfrom the combustor, the power turbine shaft would still tend to overspeed.

Accordingly, the control system as contemplated by the present inventionutilizes the guide vane actuator control 66 to shift the guide vanes120, 122 toward their FIG. 16 "reverse" position such that the gas flowfrom the engine impinges oppositely on the vanes 117, 119 of the powerturbine wheels in a manner opposing rotation of these power turbinewheels. Thus the gas flow from the engine is used to decelerate, ratherthan power, the turbine shaft 82. As a result, the power turbine shafttends to reduce in speed to the point where synchronous shifting of gearbox 38 and consequent re-engagement of drive clutch 36 may beconveniently and speedily accomplished without damage to the engine ordrive train.

More particularly, the hydromechanical portion of guide vane control 66is so arranged that upon release of accelerator lever 184 such as duringgear shifting, a very large error signal is created by the high pressurefrom the power turbine speed sensor line 420 to rotate lever 396counter-clockwise and substantially greatly increase the compression ofspring 384. Sufficient compression of spring 384 results to urge valve380 upwardly and drive piston 366 downwardly to its position illustratedin FIG. 12. This position of piston 366 corresponds to positioning theguide vanes 120, 122 in their FIG. 16 disposition. The gas flow from thecombustor is then directed by the guide vane across the turbine wheelvanes 117, 119 in opposition to the rotation thereof to decelerate thepower turbine shaft 82. Since the drive clutch 34 is disengaged duringthis gear shifting operation, the power turbine shaft 82 rather rapidlydecelerates by virtue of the opposing gas flow created by thepositioning of guide vanes 120 in their FIG. 16 position. Yet morespecifically, the arrangement of springs 406, 408 and the relativemagnitude of pressure developed in conduit 410 and 420 causes thehydromechanical portion of vane actuator control 66 to operate in themanner above described to shift the guide vanes 120 to their negative orreverse disposition illustrated in FIG. 16 and modulate guide vaneposition within zone "d" of FIG. 18 in relation to the magnitude ofN_(pt) excess, whenever the accelerator lever 184 is moved to less thana preselected accelerator lever position a*. As the speed of powerturbine shaft 82 reduces, the piston 416 begins shifting in an oppositedirection to reduce compression of spring 384 once turbine speed reducesto a preselected value. The action of piston 416 is in the preferredform capable of modulating the degree of compression of spring 384 inrelation to the magnitude of the N_(pt) error. The greater the speederror, the more the guide vanes are rotated to a "harder" brakingposition. Thus, the position of the guide vanes are maintained in areverse braking mode and are modulated through zone "d" near the maximumbraking position -95° of FIG. 18 in relation to the power turbine speederror. Once gear shifting is completed, of course, the control systemoperates through the acceleration operation discussed previously toagain increase power turbine speed.

Deceleration

A first mode of deceleration of the gas turbine engine is accomplishedby reduction in fuel flow along the deceleration schedule afforded bydeceleration window 286 of scheduling valve 62. More particularly, therelease of accelerator lever 184 causes the fuel governor 60 to severelyrestrict fuel flow through opening 178. As a consequence the minimumfuel flow to the gas turbine engine is provided through decelerationfuel line 142 and the associated deceleration window 286 of thescheduling valve. As noted previously deceleration window 286 isparticularly configured to the gas turbine engine so as to continuallyreduce fuel flow along a schedule which maintains combustion in thecombustor 98, i.e., substantially along the operating line of the gasturbine engine to maintain combustion but below the "required to runline." As noted previously, even without rotation of accelerator lever184, the solenoid 239 can be energized in particular instances togenerate a false accelerator lever signal to fuel lever 226 toaccomplish deceleration by severely restricting fuel flow.

This deceleration by limiting fuel flow is accomplished by reducing theaccelerator lever to a position at or just above a preselectedaccelerator position, a*. This accelerator position is normally justslightly above the minimum accelerator position, and generallycorresponds to the position of the accelerator lever during the"coasting" condition wherein the engine is somewhat driven by theinertia of the vehicle such as when coasting downhill. Since thisdeceleration by restricting fuel flow is acting only through governor60, it will be apparent that the guide vane control is unaffectedthereby and continues operating in the modes and conditions discussedpreviously. This is particularly true since the accelerator has beenbrought down to, but not below the preselected accelerator position a*to which the hydromechanical portion of vane actuator 66 is responsive.

Upon further rotating accelerator lever 184 below the position a* andtowards its minimum position, a second mode of deceleration or brakingof the vehicle occurs. In this mode, the movement of the acceleratorlever below the position a* causes the hydromechanical portion of guidevane actuator 66 to generate a substantially large error signal withregard to power turbine speed so as to rotate the guide vanes 120 totheir FIG. 16 reverse or "braking" position. More particularly, asdiscussed above with regard to the gear shift operation of the vehicle,this large error signal of the power turbine speed in comparison to theaccelerator lever position causes significant counter-clockwise rotationof lever 396 and consequent compression of spring 384. This drives thepiston 366 and the guide vanes toward the FIG. 16 position thereof. As aresult, the gas flow from the gas turbine engine opposes rotation of theturbine wheels 116, 118 and produces substantial tendency fordeceleration of output shaft 82. It has been found that for a gasturbine engine in the 450 to 600 horsepower class, that this reversingof the guide vanes in combination with minimum fuel flow to thecombustor as permitted by deceleration window 286 provides on the orderof 200 or more horsepower braking onto the turbine output shaft 82.

It will be noted that during this second mode of deceleration, as wellas during the gear shift operation discussed previously, that since theguide vanes are now in a reversed disposition, the logic accomplished bythe electronic control module 68 in controlling solenoid 426 to preventover temperature of T₄ or T₆ is now opposite to that required.Accordingly, the electronic control logic further includes a transducer544 which senses whenever the guide vanes pass over center as noted bythe predetermined angle B* of FIG. 18, and are in a negative incidencedisposition. This signal generated by transducer 544 energizes areversing device such as an inverter 546 which reverses the signal tothe solenoid 426. More particularly, if during this decelerationoperation with the guide vanes in the negative incidence position ofFIG. 16, there occurs an excess combustor exhaust temperature T₄ orexcess turbine exhaust temperature T₆, the signal generated by element500 to reduce the magnitude of the current signal is reversed by element546. Accordingly occurrence high T₄ or high T₆ while element 546 isenergized generates an electrical signal of increasing strength tosolenoid 426. In response, the solenoid 426 drives valve 432 in adirection increasing pressure in conduit 394 and upon shoulder 393. Thisreduces the magnitude of the biasing spring 383 and causes valve 380 tomove downwardly. In a following movement the piston 366 moves upwardlyto reduce the compression of spring 382. Thus the guide vanes 120 arereversely trimmed away from the maximum braking position shown in FIG.16 back towards the neutral position of FIG. 14. This movement of coursereduces the magnitude of power transmitted from the gas flow in opposingrotation of the guide vanes 117 to cause a consequent reduction in fuelflow as discussed previously. The reduced fuel flow then reduces themagnitude of the over temperature parameter T₄ or T₆. Such action tocontrol T₄ or T₆ will substantially only occur when fuel flow beingdelivered to the combustor is more than permitted by the decelerationschedule 286. Thus such action is more likely to occur during the"coasting" operation than during hard braking during the second mode ofdeceleration. Such is natural with operation of the engine, however,since during hard deceleration, fuel flow to the combustor is at aminimum and combustor exhaust temperature is relatively low. However,during unusual conditions, and even with the guide vanes in a negativeincidence position, the electronic control module is still operable toreturn the guide vanes toward their neutral position to tend to reduceany over temperature conditions.

Power Feedback Braking

A third mode of deceleration of the vehicle can be manually selected bythe operator. Such will normally occur when, after initiation of thefirst two modes of deceleration described above, the vehicle still isbeing driven by its own inertia at too high a speed, i.e. power turbineshaft 82 speed N_(pt) is still too high. Thus power turbine shaft speedN_(pt) may be in a range of approximately 90% of its maximum speed whilethe gas generator speed N_(gg) has been brought down to at or near itslow idle speed of approximately 50% maximum gas generator speed.

This third mode of deceleration, termed power feedback braking, ismanually selected by closing power feedback switch 466. In response theelectronic control module 68 generates signals which ultimately resultin mechanical interconnection of the gas generator shaft with the powerturbine shaft such that the inertia of the gas generator shaft isimposed upon the drive train of the vehicle to produce additionalbraking effects thereon. More particularly, upon closing switch 466, ANDgate 506 generates a signal to AND gate 504 since the accelerator levelis below a preselected point a* causing function generator 488 togenerate a signal to AND gate 506, and since the gas generator isoperating at a speed above 45% of its rated value as determined byelement 474. Element 472 develops a signal through line 480 to AND gate504 since power turbine speed is greater than gas generator speed inthis operational mode. Element 470 also notes that the effectiverelative speeds of the gas generator shaft and power turbine shaft areoutside a preselected limit, such as the plus or minus 5% noted atcomparator 470. Accordingly element 470 does not generate a signal toAND gates 502, 504. More specifically the element 470 is not necessarilycomparing the actual relative speeds of the gas generator power turbineshafts. Rather, the element is so arranged that it only generates asignal to AND gates 502, 504 whenever the relative speeds of the shafts520, 522 in the power feedback clutch 84 are within the preselectedpredetermined limits of one another. Thus the comparator 468 willcompensate, as required, for differences in the actual speeds of the gasgenerator and power turbine shaft, as well as the gear ratios of the tworespective drive trains 78 and 80 associated with the two shafts 502,522 of the feedback clutch 84.

Because of the difference between N_(pt) and N_(gg), no signal fromelement 470 is transmitted to either AND gate 502 or 504. As notedschematically by the circle associated with the input from element 470to AND gate 504, that input is inverted and AND gate 504 is noweffective to generate an output signal since no signal is coming fromelement 470, and since signals are being received from AND gate 506 andelement 472. The output signal from AND gate 504 accomplishes twofunctions. First, a signal of 50% N_(gg) magnitude is generated infunction generator 566 and added to the constant 50% bias signal ofsummer 570. The resulting signal is equivalent to a 100% N_(gg) speedcommand. Secondly, the output from AND gate 504 passes through OR gate562 to produce a signal to solenoid 257. This signal is of sufficientmagnitude to shift solenoid 257 to its FIG. 6D position clearing opening178 for substantial fuel flow to the combustor. It will be apparent thatfull energization of solenoid 257 to its FIG. 6D position is essentiallya false accelerator lever signal to the fuel lever 226 causing lever 226to rotate to a position normally caused by depressing accelerating lever184 to its maximum flow position. Secondly, the signal from summer 570is also an input to element 497 such that an artificial full throttlesignal is generated which overrides the energization signal which ismaintaining the guide vanes in their FIG. 16 braking position during thesecond mode of deceleration discussed previously. The energization ofthe guide vane solenoid 426 causes increase of pressure in conduit 394allowing the springs 382-385 to shift the piston 366 and the associatedguide vanes toward their FIG. 14 "neutral" position.

Accordingly, it will be seen that the signal from AND gate 504 producesan acceleration signal to the engine, placing the guide vanes 120, 122in their neutral position such that maximum pressure ratio is developedacross the gas generator turbine 102, and at the same time fuel flow tothe combustor 98 has been greatly increased. In response, the gasgenerator section begins increasing in speed rapidly toward a value suchthat the speed of shaft 522 of the feedback clutch approaches the speedof its other shaft 520.

Once the power turbine and gas generator shaft speeds are appropriatelymatched such that the two shafts 520, 522 of the feedback clutch arewithin the preselected limits determined by element 470 of theelectronic control module, electronic control module develops a positivesignal to both AND gates 502, 504. This positive signal immediatelystops the output signal from AND gate 504 to de-energize theproportional solenoid 257 of the fuel governor and again reduce fuelflow back toward a minimum value, and at the same time stops theoverride signal to element 500 such that the guide vane 120, 122 areagain shifted back to their FIG. 16 braking disposition in accord withthe operation discussed above with respect with the second mode ofdeceleration.

The logic element AND gate 502 now develops a positive output signal tooperate to driver 516 and energize clutch actuator solenoid valve 518.In response the clutch 84 becomes engaged to mechanically interlock theshafts 520 and 522 as well as the gas generator and power turbine shafts76, 82. Incorporation of the logic element 470 in the electronic controlmodule, in addition to the other functions described previously, alsoassures that since the two shafts 520, 522 are in near synchronousspeed, relatively small torque miss-match across the plates 524, 526 ofthe clutch is experienced. Accordingly, the size of clutch 84 can berelatively small. Thus it will be seen that the electronic controlmodule 68 operates automatically first to increase gas generator speedto essentially match power turbine speed, and then to automaticallyreturn the guide vanes to their FIG. 16 braking disposition at the sametime as clutch 84 is engaged.

This interconnection of the gas turbine engine drive train with the gasgenerator shaft 76 causes the rotational inertia of gas generator 76 toassist in decelerating the vehicle. It has been found that for a 450 to600 horsepower class engine described, this power feedback braking modeadds in the neighborhood of 200 to 250 horsepower braking in addition tothe 200 horsepower braking effects produced by the positioning of guidevane 120, 122 in their FIG. 16 position. Because the fuel governor isagain severely restricting flow through orifice 178, the fuel flow iscontrolled by deceleration window 286 permitting the gas generatorsection to decelerate while maintaining the combustion process incombustor 98. Thus reduction of fuel flow provides the decelerationeffect of the rotational inertia of the gas generator upon the drivetrain of the vehicle.

It will be apparent from the foregoing that the present inventionprovides substantial braking for deceleration purposes while stillutilizing the optimum operating characteristics of a free turbine typeof a gas turbine engine with the gas generator section mechanicallyinterconnected with the power turbine section only in a specificinstance of a manually selected "severe" third mode type of decelerationoperation. Throughout all deceleration modes and engine operation, acontinuous combustion process is maintained in the combustor. Thussubstantial deceleration occurs without extinguishing the combustionprocess therein.

This power feedback braking operation can be deactivated in severalways: manually by opening switch 466 to stop the output signal from ANDgate 506; providing a NOT signal to turn off driver 516 and solenoid 518to disengage clutch 84. Furthermore, if the manual switch is not openedand the engine continues to decelerate, element 474 also operates todeactivate the power feedback operation whenever gas generator speedN_(gg) reduces to a value below 45% of its maximum rate of speed. Also,depression of the accelerator to a value of above a* also deactivatesthe power feedback operation by stopping an output signal from AND gate506.

From the foregoing it will now be apparent that the present inventionprovides an improved cycle of operation for a gas turbine enginepeculiarly adapted for operating a ground vehicle in a safe, familiarmanner while still retaining the inherent benefits of a gas turbineengine. More specifically, by utilization of a free turbine type enginegreater adaptability and variability of engine operation is provided. Atthe same time the engine can operate throughout its entire operatingcycle while maintaining a continuous combustion process within thecombustor 98. This avoids various problems of operation and service lifeassociated with repeated start and stop of the combustion process. Thenovel cycle contemplates a utilization of a combustor 98 having chokednozzles 102 to provide a variable pressure within the combustor as thespeed of the gas generator section varies. Gas generator section speedis normally trimmed to a preselected value relative to the position ofthe accelerator lever 184, while the guide vanes 120, 122 operate totrim the turbine inlet temperature T₄ to a preselected substantiallyconstant value to maintain high engine operational efficiency. Further,the guide vane control operates indirectly to alter the fuel flowthrough fuel governor 60 by altering the speed of the gas generatorsection such that the various controls are operable in an integralmanner without counteracting one another. At the same time a trim ofpower turbine shaft speed N_(pt) is provided by the guide vane control66.

Furthermore it will be seen that the present invention provides the gasturbine engine peculiarly adapted for driving a ground vehicle in thatresponsive acceleration similar to that produced by an internalcombustion engine is provided by both the automatic high idle operationas well as by the manner of acceleration of the gas turbine engine. Suchis accomplished by first altering the work split to develop maximumpower to the gas generator section. The scheduling valve control 62 thenacts in regenerative fashion to increase fuel flow to the combustor insuch a manner that gas generator speed is increased while maintaining asubstantially constant maximum turbine inlet temperature T₄ therebyproducing maximum acceleration without over heating the engine. Yet thescheduling valve also limits T₆ during the initial portion ofacceleration when turbine "stalling" conditions are prevalent.Acceleration is then completed once substantial acceleration of the gasgenerator section is accomplished, by re-altering the power split todevelop greater power across the power turbine wheels 116, 118.

It is further noted that the present invention provides an improvedmethod and apparatus for decelerating the vehicle in a three stage typeof operation by first reducing fuel flow, then by placing the guidevanes in the braking mode, and then by manually selecting the powerfeedback operation.

The primary operating elements of the fuel governor 60, scheduling valve62, and guide vane control 66 are hydromechanical in nature. This, inconjunction with the operation of solenoid 426 of the guide vane controlwhich is normally energized, provides an engine and control systempeculiarly adapted to provide safe engine operation in the event ofvarious failure modes. More particularly, in the event of complete lossof electrical power to the electronic control module 68, the mechanicalportion of fuel governor 60 continues to adjust fuel flow in relation tothat selected by accelerator lever 184. Scheduling valve 62 is in no wayaffected by such electrical failure and is capable of controllingacceleration and/or deceleration to both prevent over heating of theengine during acceleration as well as to maintain combustion duringdeceleration. The hydromechanical portion of the vane actuator controlwill still be operable in the event of electrical failure and capable ofadjusting the guide vanes as appropriate to maintain functional engineoperation. Upon electrical failure the solenoid 426 of the guide vanecontrol becomes de-energized causing loss of pressure upon face 393 ofthe control piston 392. However, the speed control afforded by lever 396is still maintained and the guide vanes can be appropriately positionedto maintain functional engine operation during this failure of theelectrical system. Thus, while certain desirable features of the enginecontrol will be lost in the event of electrical failure, the engine canstill function properly with appropriate acceleration and decelerationso that the vehicle may still be operated in a safe manner even thoughat a possible loss of operational efficiency and loss of the ability toprovide power feedback braking.

From the foregoing it will be apparent that the present inventionprovides an improved method of automatically setting and resetting theidle of the gas generator section so that the engine is highlyresponsive in developing an increase in output power such as whencontemplating acceleration of the vehicle. Further the present inventionprovides an improved method of controlling fuel flow hydromechanicallyin relation to gas generator speed, as well as overriding normal speedcontrol operation of the fuel governor to increase or decrease fuel flowin response to occurrence of various conditions which energize either ofthe solenoids 239, 257. Further the present invention provides animproved method for controlling fuel flow to the combustor duringacceleration such that constant turbine inlet temperature T₄ ismaintained throughout, while also controlling fuel flow duringdeceleration to avoid extinguishing the combustion process within acombustor. The invention further contemplates an improved method ofcontrolling guide vane position in such an engine both byhydromechanical operation to control speed of a rotor such as turbinewheels 116, 118, and by electrical override operation dependent upon theamount of energization of the proportional solenoid 426.

The foregoing has described a preferred embodiment of the invention insufficient detail that those skilled in the art may make and use it.However, this detailed description should be considered exemplary innature and not as limiting to the scope and spirit of the presentinvention as set forth in the appended claims.

Having described the invention with sufficient clarity that thoseskilled in the art may make and use it, what is claimed as new anddesired to be secured by Letters Patent is:
 1. A method of altering theincidence of motive gas flow onto a power turbine section of a freeturbine type gas turbine engine, comprising the steps of:mechanicallysensing power turbine section speed; mechanically sensing the incidenceof gas flow onto the power turbine section; generating a mechanicalinput signal indicative of a desired speed for the power turbinesection; comparing the sensed speed with the desired speed input signaland generating a mechanical speed error signal indicative of thedifference between the sensed and desires speeds; comparing the sensedincidence of gas flow with the speed error signal and generating amechanical output signal indicative of the difference between the sensedincidence and the speed error signal; and adjusting the incidence of gasflow onto the power turbine section in relation to said mechanicaloutput signal in a manner minimizing said speed error signal.
 2. Amethod as set forth in claim 1, further including the steps ofgenerating an electrical signal in relation to a parameter of engineoperation other than power turbine section speed, and transforming saidelectrical signal into a corresponding mechanical override signalcapable of overriding said output signal to adjust the incidence of gasflow in relation to said parameter regardless of said speed errorsignal.
 3. A method as set forth in claim 2, said transforming stepbeing accomplished in a manner whereby upon loss of electrical power,gas flow incidence is adjusted in relation to said mechanical outputsignal.
 4. A control for adjusting the position of guide vanes to alterthe incidence of motive gas flow onto a rotor of a gas turbine engine,comprising:an output member operably engaging said guide vanes andshiftable to adjust guide vane position; an input member movable toeffect shifting of said output member; means for sensing the speed ofsaid rotor; throttle means for developing a mechanical signal indicativeof a desired speed for said rotor; mechanical comparing means responsiveto said speed sensing means and said mechanical signal of said throttlemeans, said comparing means operably associated with said input memberfor exerting a mechanical input force thereon indicative of thedifference between said desired speed and said sensed speed; andmechanical feedback means operably engaging said input and outputmembers for exerting a mechanical feedback force on said input memberopposing said mechanical input force in relation to the position of saidmember, whereby said feedback and input forces move said input member toeffect corresponding shifting of said output member and guide vanes tominimize said difference between the desired and sensed rotor speeds. 5.A control as set forth in claim 4, wherein said engine is of the freeturbine type having a gas generator section and a pwoer turbine sectionrotatable independently of said gas generator section, said rotor beinga turbine in said power turbine section.
 6. A control as set forth inclaim 5, further including an electronic control for sensing a parameterof engine operation other than said rotor speed and for generating anelectrical signal indicative of said parameter; and transducer meansresponsive to said electrical signal and operably associated with saidinput member for exerting a mechanical override force on said inputmember in relation to said electrical signal, said override forcecapable of overriding said input force to move said input member inrelation to said parameter regardless of said difference between thedesired and sensed speeds.
 7. A control as set forth in claim 6, whereinsaid transducer means is arranged and configured whereby upon loss ofelectrical power to said electronic control, said input member is movedby said feedback and input forces.
 8. A control as set forth in claim 7,wherein said parameter is the temperature of gas flow exhausting fromsaid power turbine section.
 9. A control as set forth in claim 7,wherein said parameter is a preselected condition of engine operation.10. A control as set forth in claim 7, wherein said gas generatorsection includes a combustor for generating said motive gas flow, saidparameter being the temperature of motive gas flow exhausting from saidcombustor.
 11. A control as set forth in claim 10, wherein saidtransducer means is operable to generate an override force of decreasingmagnitude with increasing temperature of exhaust gas flow from saidcombustor, said override force operable to control the position of saidguide vanes to limit said combustor exhaust temperature.
 12. A controlas set forth in claim 11, wherein said transducer means is arrangedwhereby the maximum value of said override force occurs when saidelectrical signal is at a minimum value.
 13. A control as set forth inclaim 12, including an adjustable stop; compressible biasing meansextending between said stop and said input member for exerting saidoverride force on said input member, said transducer means operable tomove said stop in a direction reducing the compression of said biasingmeans and the magnitude of said override force as the strength of saidelectrical signal increases; and a biasing element exerting a biasingforce on said stop urging the latter in an opposite direction increasingthe compression of said biasing means and the magnitude of said overrideforce.
 14. A control as set forth in claim 13, further including asource of pressurized fluid, said adjustable stop having a surface areaexposed to a fluid chamber, said transducer means comprising a solenoidhaving a valving member positioned in response to said electrical signalto control fluid communication of said source with said chamber to alterpressure of fluid in said chamber as a function of said electricalsignal.
 15. A control as set forth in claim 14, wherein said transducermeans includes a spring urging said valving member in opposition to saidelectrical signal and toward a position developing maximum fluidpressure in said chamber.
 16. A control as set forth in claim 15,wherein said valving member is operable to reduce fluid pressure in saidchamber as said combustor exhaust temperature increases.
 17. A controlas set forth in claim 4, further including a source of pressurizedfluid, a housing having inlet and outlet ports and an internal cylinder,said output member including a double acting piston traversing saidcylinder to divide the latter into opposed variable volume chambers, andlinkage operably interconnecting said piston and said guide vanes.
 18. Acontrol as set forth in claim 17, wherein said output member comprises afour-way valve shiftable in a bore in said housing to control fluidcommunication of said inlet and output ports with said opposed variablevolume chambers.
 19. A control as set forth in claim 18, wherein saidfeedback means comprises a compressible biasing spring extending betweensaid piston and said four-way valve exerting said feedback force to urgesaid valve in a first direction.
 20. A control as set forth in claim 19,wherein said comparing means includes a second spring exerting saidinput force to urge said four-way valve in a second, opposite direction.21. In a free turbine type gas turbine engine having a power turbinesection driven by a motive gas flow developed by a gas generator sectionof the engine:variably positionable guide vanes disposed in said gasflow for altering the incidence thereof upon said power turbine section;a source of pressurized fluid; a housing having a fluid exhaust port, aninlet port communicating with said source, an internal cylinder, and aninternal bore communicating with said inlet and outlet ports; a pistonmovable in said cylinder and dividing the latter into opposed fluidchambers communicating with said bore at spaced locations therealong;linkage operably interconnecting said piston and said guide vaneswhereby said guide vanes are positioned in relation to the position ofsaid piston; a four-way valve movable within said bore to controlcommunication of said opposed chambers with said inlet and outlet ports;a first feedback spring means extending between said piston and saidfour-way valve to urge the latter in a first direction in relation tothe position of said piston; a first stop movably mounted in saidhousing; second input spring means extending between said first stop andsaid four-way valve to urge the latter in a second opposite direction inrelation to the position of said first stop; and input means foradjusting the position of said first stop.
 22. A gas turbine engine asset forth in claim 21, wherein said input means includes a second fluidoperated piston disposed in a second bore in said housing and operablyengaging said first stop, and manual signal means for exerting a firstfluid force on said second piston urging the latter to move in adirection reducing the urgings of said input spring means on saidfour-way valve.
 23. A gas turbine engine as set forth in claim 22,wherein said manual signal means includes a manually positionablethrottle, and a second valve for metering fluid flow to one side of saidsecond piston to vary the fluid pressure exerted on said one side inrelation to throttle position.
 24. A gas turbine engine as set forth inclaim 23, wherein said input means further includes speed sensing meansassociated with said power turbine section for exerting a second fluidforce on an opposite side of said second piston opposing said firstfluid force, the magnitude of said second fluid force being indicativeof the speed of said power turbine section.
 25. A gas turbine engine asset forth in claim 24, wherein said speed sensing means is operable tovary the fluid pressure exerted on said opposite side of the piston inrelation to said speed of the power turbine section.
 26. A control asset forth in claim 25, further including an electronic control forsensing a parameter of engine operation other than said power turbinespeed and for generating an electrical signal indicative of saidparameter; and transducer means responsive to said electrical signal andoperably associated with said four-way valve for exerting a mechanicaloverride force on said four-way valve in relation to said electricalsignal, said override force capable of overriding said input springmeans to move said four way valve in relation to said parameter.
 27. Acontrol as set forth in claim 26, wherein said transducer means isarranged and configured whereby upon loss of electrical power to saidelectronic control, said four-way valve is moved by said feedback springand input spring means.
 28. A control as set forth in claim 27, whereinsaid transducer means is arranged whereby the maximum value of saidoverride force occurs when said electrical signal is at a minimum value.29. A control as set forth in claim 28, including a second adjustablestop; a third spring extending between said second stop and saidfour-way valve for exerting said override force on said four-way valve,said transducer means operable to move said second stop in a directionreducing the compression of said third spring and the magnitude of saidoverride force as the strength of said electrical signal increases; anda fourth spring exerting a biasing force on said second stop urging thelatter in an opposite direction increasing the compression of said thirdspring and the magnitude of said override force.
 30. A control as setforth in claim 29, wherein said second adjustable stop comprises astepped piston in said first mentioned bore having a surface areaexposed to another fluid chamber, said transducer means comprising asolenoid having a third valving member positioned in response to saidelectrical signal to control fluid communication of said source withsaid another chamber to alter pressure of fluid in said another chamberas a function of said electrical signal.
 31. A control as set forth inclaim 30, wherein said transducer means includes a fifth spring urgingsaid third valving member in opposition to said electrical signal andtoward a position developing maximum fluid pressure in said anotherchamber.
 32. A control as set forth in claim 31, wherein said thirdvalving member is operable to reduce fluid pressure in said anotherchamber as said electrical signal increases in strength.
 33. A controlas set forth in claim 30, wherein said first stop includes a plungeroperably engaged by said second piston, said plunger extending through abore in said second stop to be movable independently thereof, saidsecond and third springs being concentrically arranged and contactingthe same end of said four-way valve.